2014 ASHRAE Handbook -- Refrigeration (SI).pdf

2014 ASHRAE® HANDBOOK

REFRIGERATION

SI Edition

ASHRAE, 1791 Tullie Circle, N.E., Atlanta, GA 30329 www.ashrae.org

© 2014 ASHRAE. All rights reserved. DEDICATED TO THE ADVANCEMENT OF THE PROFESSION AND ITS ALLIED INDUSTRIES

No part of this publication may be reproduced without permission in writing from ASHRAE, except by a reviewer who may quote brief passages or reproduce illustrations in a review with appropriate credit; nor may any part of this book be reproduced, stored in a retrieval system, or transmitted in any way or by any means—electronic, photocopying, recording, or other—without permission in writing from ASHRAE. Requests for permission should be submitted at www.ashrae.org/permissions. Volunteer members of ASHRAE Technical Committees and others compiled the information in this handbook, and it is generally reviewed and updated every four years. Comments, criticisms, and suggestions regarding the subject matter are invited. Any errors or omissions in the data should be brought to the attention of the Editor. Additions and corrections to Handbook volumes in print will be published in the Handbook published the year following their verification and, as soon as verified, on the ASHRAE Internet web site. DISCLAIMER ASHRAE has compiled this publication with care, but ASHRAE has not investigated, and ASHRAE expressly disclaims any duty to investigate, any product, service, process, procedure, design, or the like that may be described herein. The appearance of any technical data or editorial material in this publication does not constitute endorsement, warranty, or guaranty by ASHRAE of any product, service, process, procedure, design, or the like. ASHRAE does not warrant that the information in this publication is free of errors. The entire risk of the use of any information in this publication is assumed by the user. ISBN 978-1-936504-72-5 ISSN 1930-7217

The paper for this book is both acid- and elemental-chlorine-free and was manufactured with pulp obtained from sources using sustainable forestry practices.

ASHRAE TECHNICAL COMMITTEES, TASK GROUPS, AND TECHNICAL RESOURCE GROUPS SECTION 1.0—FUNDAMENTALS AND GENERAL 1.1 Thermodynamics and Psychrometrics 1.2 Instruments and Measurements 1.3 Heat Transfer and Fluid Flow 1.4 Control Theory and Application 1.5 Computer Applications 1.6 Terminology 1.7 Business, Management, and General Legal Education 1.8 Mechanical Systems Insulation 1.9 Electrical Systems 1.10 Cogeneration Systems 1.11 Electric Motors and Motor Control 1.12 Moisture Management in Buildings TG1 Optimization SECTION 2.0—ENVIRONMENTAL QUALITY 2.1 Physiology and Human Environment 2.2 Plant and Animal Environment 2.3 Gaseous Air Contaminants and Gas Contaminant Removal Equipment 2.4 Particulate Air Contaminants and Particulate Contaminant Removal Equipment 2.5 Global Climate Change 2.6 Sound and Vibration Control 2.7 Seismic and Wind Resistant Design 2.8 Building Environmental Impacts and Sustainability 2.9 Ultraviolet Air and Surface Treatment TG2 Heating, Ventilation, and Air-Conditioning Security (HVAC) SECTION 3.0—MATERIALS AND PROCESSES 3.1 Refrigerants and Secondary Coolants 3.2 Refrigerant System Chemistry 3.3 Refrigerant Contaminant Control 3.4 Lubrication 3.6 Water Treatment 3.8 Refrigerant Containment SECTION 4.0—LOAD CALCULATIONS AND ENERGY REQUIREMENTS 4.1 Load Calculation Data and Procedures 4.2 Climatic Information 4.3 Ventilation Requirements and Infiltration 4.4 Building Materials and Building Envelope Performance 4.5 Fenestration 4.7 Energy Calculations 4.10 Indoor Environmental Modeling TRG4 Indoor Air Quality Procedure Development SECTION 5.0—VENTILATION AND AIR DISTRIBUTION 5.1 Fans 5.2 Duct Design 5.3 Room Air Distribution 5.4 Industrial Process Air Cleaning (Air Pollution Control) 5.5 Air-to-Air Energy Recovery 5.6 Control of Fire and Smoke 5.7 Evaporative Cooling 5.8 Industrial Ventilation 5.9 Enclosed Vehicular Facilities 5.10 Kitchen Ventilation 5.11 Humidifying Equipment SECTION 6.0—HEATING EQUIPMENT, HEATING AND COOLING SYSTEMS AND APPLICATIONS 6.1 Hydronic and Steam Equipment and Systems 6.2 District Energy 6.3 Central Forced Air Heating and Cooling Systems 6.5 Radiant Heating and Cooling 6.6 Service Water Heating Systems

6.7 6.8 6.9 6.10

Solar Energy Utilization Geothermal Heat Pump and Energy Recovery Applications Thermal Storage Fuels and Combustion

SECTION 7.0—BUILDING PERFORMANCE 7.1 Integrated Building Design 7.2 HVAC&R Construction and Design Build Technologies 7.3 Operation and Maintenance Management 7.4 Exergy Analysis for Sustainable Buildings (EXER) 7.5 Smart Building Systems 7.6 Building Energy Performance 7.7 Testing and Balancing 7.8 Owning and Operating Costs 7.9 Building Commissioning SECTION 8.0—AIR-CONDITIONING AND REFRIGERATION SYSTEM COMPONENTS 8.1 Positive Displacement Compressors 8.2 Centrifugal Machines 8.3 Absorption and Heat Operated Machines 8.4 Air-to-Refrigerant Heat Transfer Equipment 8.5 Liquid-to-Refrigerant Heat Exchangers 8.6 Cooling Towers and Evaporative Condensers 8.7 Variable Refrigerant Flow (VRF) 8.8 Refrigerant System Controls and Accessories 8.9 Residential Refrigerators and Food Freezers 8.10 Mechanical Dehumidification Equipment and Heat Pipes 8.11 Unitary and Room Air Conditioners and Heat Pumps 8.12 Desiccant Dehumidification Equipment and Components SECTION 9.0—BUILDING APPLICATIONS 9.1 Large Building Air-Conditioning Systems 9.2 Industrial Air Conditioning 9.3 Transportation Air Conditioning 9.4 Justice Facilities 9.6 Healthcare Facilities 9.7 Educational Facilities 9.8 Large Building Air-Conditioning Applications 9.9 Mission Critical Facilities, Data Centers, Technology Spaces and Electronic Equipment 9.10 Laboratory Systems 9.11 Clean Spaces 9.12 Tall Buildings SECTION 10.0—REFRIGERATION SYSTEMS 10.1 Custom Engineered Refrigeration Systems 10.2 Automatic Icemaking Plants and Skating Rinks 10.3 Refrigerant Piping, Controls and Accessories 10.5 Refrigerated Distribution and Storage Facilities 10.6 Transport Refrigeration 10.7 Commercial Food and Beverage Refrigeration Equipment 10.8 Refrigeration Load Calculations SECTION MTG—MULTIDISCIPLINARY TASK GROUPS MTG.BD Building Dampness MTG.BIM Building Information Modeling MTG.CCDG Cold Climate Design Guide MTG.EAS Energy-Efficient Air Handling Systems for NonResidential Buildings MTG.ET Energy Targets MTG.HCDG Hot Climate Design Guide MTG.LowGWP Lower Global Warming Potential Alternative Refrigerants

ASHRAE Research ASHRAE is the world’s foremost technical society in the fields of heating, ventilation, air conditioning, and refrigeration. Its members worldwide are individuals who share ideas, identify needs, support research, and write the industry’s standards for testing and practice. The result is that engineers are better able to keep indoor environments safe and productive while protecting and preserving the outdoors for generations to come. One of the ways that ASHRAE supports its members’ and industry’s need for information is through ASHRAE Research. Thousands of individuals and companies support ASHRAE Research annually, enabling

ASHRAE to report new data about material properties and building physics and to promote the application of innovative technologies. Chapters in the ASHRAE Handbook are updated through the experience of members of ASHRAE Technical Committees and through results of ASHRAE Research reported at ASHRAE conferences and published in ASHRAE special publications and in ASHRAE Transactions. For information about ASHRAE Research or to become a member, contact ASHRAE, 1791 Tullie Circle, Atlanta, GA 30329; telephone: 404-636-8400; www.ashrae.org.

Preface The 2014 ASHRAE Handbook—Refrigeration covers the refrigeration equipment and systems for applications other than human comfort. This volume includes data and guidance on cooling, freezing, and storing food; industrial and medical applications of refrigeration; and low-temperature refrigeration. An accompanying CD-ROM contains all the volume’s chapters in both I-P and SI units. Some of this volume’s revisions are described as follows: • Chapter 1, Halocarbon Refrigeration Systems, has three new sections to address issues involving the Montreal Protocol and the phaseout of halocarbons. It also has a new introduction, plus updates to sections on Applications and System Safety. • Chapter 2, Ammonia Refrigeration Systems, has been extensively reorganized and updated for current practice. • Chapter 6, Refrigerant System Chemistry, has new sections on additives and process chemicals. • Chapter 7, Control of Moisture and Other Contaminants in Refrigerant Systems, has added moisture isotherm data for refrigerants R-290 and R-600a. It also contains a new section on system sampling in conjunction with retrofits, troubleshooting, or routine maintenance. • Chapter 10, Insulation Systems for Refrigerant Piping, has revised insulation table values to comply with ASTM Standard C680-10. • Chapter 12, Lubricants in Refrigerant Systems, has expanded content on hydrofluorocarbons (HFCs) and new guidance on retrofits. • Chapter 15, Retail Food Store Refrigeration and Equipment, has updates to sections on multiplex compressor racks, secondary and CO2 systems, gas defrost, liquid subcooling, and heat reclaim. • Chapter 17, Household Refrigerators and Freezers, has updates on LED lighting in cabinets.

• Chapter 24, Refrigerated-Facility Loads, includes new content on packaging loads from moisture, updated motor heat gain rates, and a new example of a complete facility load calculation. • Chapter 25, Cargo Containers, Rail Cars, Trailers, and Trucks, updated throughout, has a major revision to the section on Equipment. • Chapter 27, Air Transport, has major revisions to the extensive section on Galley Refrigeration. • Chapter 51, Codes and Standards, has been updated to list current versions of selected publications from ASHRAE and others. Publications are listed by topic, and full contact information for publishing organizations is included. This volume is published, as a bound print volume and in electronic format on CD-ROM and online, in two editions: one using inch-pound (I-P) units of measurement, the other using the International System of Units (SI). Corrections to the 2011, 2012, and 2013 Handbook volumes can be found on the ASHRAE web site at http://www.ashrae.org and in the Additions and Corrections section of this volume. Corrections for this volume will be listed in subsequent volumes and on the ASHRAE web site. Reader comments are enthusiastically invited. To suggest improvements for a chapter, please comment using the form on the ASHRAE web site or, using the cutout page(s) at the end of this volume’s index, write to Handbook Editor, ASHRAE, 1791 Tullie Circle, Atlanta, GA 30329, or fax 678-539-2187, or e-mail [email protected].

Mark S. Owen Editor

ASHRAE Research ASHRAE is the world’s foremost technical society in the fields of heating, ventilation, air conditioning, and refrigeration. Its members worldwide are individuals who share ideas, identify needs, support research, and write the industry’s standards for testing and practice. The result is that engineers are better able to keep indoor environments safe and productive while protecting and preserving the outdoors for generations to come. One of the ways that ASHRAE supports its members’ and industry’s need for information is through ASHRAE Research. Thousands of individuals and companies support ASHRAE Research annually, enabling

ASHRAE to report new data about material properties and building physics and to promote the application of innovative technologies. Chapters in the ASHRAE Handbook are updated through the experience of members of ASHRAE Technical Committees and through results of ASHRAE Research reported at ASHRAE conferences and published in ASHRAE special publications and in ASHRAE Transactions. For information about ASHRAE Research or to become a member, contact ASHRAE, 1791 Tullie Circle, Atlanta, GA 30329; telephone: 404-636-8400; www.ashrae.org.

Preface The 2014 ASHRAE Handbook—Refrigeration covers the refrigeration equipment and systems for applications other than human comfort. This volume includes data and guidance on cooling, freezing, and storing food; industrial and medical applications of refrigeration; and low-temperature refrigeration. An accompanying CD-ROM contains all the volume’s chapters in both I-P and SI units. Some of this volume’s revisions are described as follows: • Chapter 1, Halocarbon Refrigeration Systems, has three new sections to address issues involving the Montreal Protocol and the phaseout of halocarbons. It also has a new introduction, plus updates to sections on Applications and System Safety. • Chapter 2, Ammonia Refrigeration Systems, has been extensively reorganized and updated for current practice. • Chapter 6, Refrigerant System Chemistry, has new sections on additives and process chemicals. • Chapter 7, Control of Moisture and Other Contaminants in Refrigerant Systems, has added moisture isotherm data for refrigerants R-290 and R-600a. It also contains a new section on system sampling in conjunction with retrofits, troubleshooting, or routine maintenance. • Chapter 10, Insulation Systems for Refrigerant Piping, has revised insulation table values to comply with ASTM Standard C680-10. • Chapter 12, Lubricants in Refrigerant Systems, has expanded content on hydrofluorocarbons (HFCs) and new guidance on retrofits. • Chapter 15, Retail Food Store Refrigeration and Equipment, has updates to sections on multiplex compressor racks, secondary and CO2 systems, gas defrost, liquid subcooling, and heat reclaim. • Chapter 17, Household Refrigerators and Freezers, has updates on LED lighting in cabinets.

• Chapter 24, Refrigerated-Facility Loads, includes new content on packaging loads from moisture, updated motor heat gain rates, and a new example of a complete facility load calculation. • Chapter 25, Cargo Containers, Rail Cars, Trailers, and Trucks, updated throughout, has a major revision to the section on Equipment. • Chapter 27, Air Transport, has major revisions to the extensive section on Galley Refrigeration. • Chapter 51, Codes and Standards, has been updated to list current versions of selected publications from ASHRAE and others. Publications are listed by topic, and full contact information for publishing organizations is included. This volume is published, as a bound print volume and in electronic format on CD-ROM and online, in two editions: one using inch-pound (I-P) units of measurement, the other using the International System of Units (SI). Corrections to the 2011, 2012, and 2013 Handbook volumes can be found on the ASHRAE web site at http://www.ashrae.org and in the Additions and Corrections section of this volume. Corrections for this volume will be listed in subsequent volumes and on the ASHRAE web site. Reader comments are enthusiastically invited. To suggest improvements for a chapter, please comment using the form on the ASHRAE web site or, using the cutout page(s) at the end of this volume’s index, write to Handbook Editor, ASHRAE, 1791 Tullie Circle, Atlanta, GA 30329, or fax 678-539-2187, or e-mail [email protected].

Mark S. Owen Editor

CONTENTS Contributors ASHRAE Technical Committees, Task Groups, and Technical Resource Groups ASHRAE Research: Improving the Quality of Life Preface SYSTEMS AND PRACTICES Chapter

1. 2. 3. 4. 5. 6. 7.

Halocarbon Refrigeration Systems (TC 10.3, Refrigerant Piping, Controls and Accessories) Ammonia Refrigeration Systems (TC 10.3) Carbon Dioxide Refrigeration Systems (TC 10.3) Liquid Overfeed Systems (TC 10.1, Custom Engineered Refrigeration Systems) Component Balancing in Refrigeration Systems (TC 10.1) Refrigerant System Chemistry (TC 3.2, Refrigerant System Chemistry) Control of Moisture and Other Contaminants in Refrigerant Systems (TC 3.3, Refrigerant Contaminant Control) 8. Equipment and System Dehydrating, Charging, and Testing (TC 8.1, Positive Displacement Compressors) 9. Refrigerant Containment, Recovery, Recycling, and Reclamation (TC 3.8, Refrigerant Containment)

COMPONENTS AND EQUIPMENT Chapter

10. 11. 12. 13. 14. 15.

Insulation Systems for Refrigerant Piping (TC 10.3) Refrigerant Control Devices (TC 8.8, Refrigerant System Controls and Accessories) Lubricants in Refrigerant Systems (TC 3.4, Lubrication) Secondary Coolants in Refrigeration Systems (TC 10.1) Forced-Circulation Air Coolers (TC 8.4, Air-to-Refrigerant Heat Transfer Equipment) Retail Food Store Refrigeration and Equipment (TC 10.7, Commercial Food and Beverage Refrigeration Equipment) 16. Food Service and General Commercial Refrigeration Equipment (TC 10.7) 17. Household Refrigerators and Freezers (TC 8.9, Residential Refrigerators and Food Freezers) 18. Absorption Equipment (TC 8.3, Absorption and Heat Operated Machines)

FOOD COOLING AND STORAGE Chapter

19. 20. 21. 22. 23. 24.

Thermal Properties of Foods (TC 10.5, Refrigerated Distribution and Storage Facilities) Cooling and Freezing Times of Foods (TC 10.5) Commodity Storage Requirements (TC 10.5) Food Microbiology and Refrigeration (TC 10.5) Refrigerated-Facility Design (TC 10.5) Refrigerated-Facility Loads (TC 10.8, Refrigeration Load Calculations)

REFRIGERATED TRANSPORT Chapter

25. Cargo Containers, Rail Cars, Trailers, and Trucks (TC 10.6, Transport Refrigeration)

26. Marine Refrigeration (TC 10.6) 27. Air Transport (TC 10.6)

FOOD, BEVERAGE, AND FLORAL APPLICATIONS Chapter

28. 29. 30. 31. 32. 33. 34. 35. 36. 37. 38. 39. 40. 41. 42.

Methods of Precooling Fruits, Vegetables, and Cut Flowers (TC 10.5) Industrial Food-Freezing Systems (TC 10.5) Meat Products (TC 10.5) Poultry Products (TC 10.5) Fishery Products (TC 10.5) Dairy Products (TC 10.5) Eggs and Egg Products (TC 10.5) Deciduous Tree and Vine Fruit (TC 10.5) Citrus Fruit, Bananas, and Subtropical Fruit (TC 10.5) Vegetables (TC 10.5) Fruit Juice Concentrates and Chilled Juice Products (TC 10.5) Beverages (TC 10.5) Processed, Precooked, and Prepared Foods (TC 10.5) Bakery Products (TC 10.5) Chocolates, Candies, Nuts, Dried Fruits, and Dried Vegetables (TC 10.5)

INDUSTRIAL APPLICATIONS Chapter

43. 44. 45. 46.

Ice Manufacture (TC 10.2, Automatic Icemaking Plants and Skating Rinks) Ice Rinks (TC 10.2) Concrete Dams and Subsurface Soils (TC 10.1) Refrigeration in the Chemical Industry (TC 10.1)

LOW-TEMPERATURE APPLICATIONS Chapter

47. Cryogenics (TC 10.1) 48. Ultralow-Temperature Refrigeration (TC 10.1) 49. Biomedical Applications of Cryogenic Refrigeration (TC 10.1)

GENERAL Chapter

50. Terminology of Refrigeration (TC 10.1) 51. Codes and Standards

ADDITIONS AND CORRECTIONS INDEX Composite index to the 2011 HVAC Applications, 2012 HVAC Systems and Equipment, 2013 Fundamentals, and 2014 Refrigeration volumes

Comment Pages

CHAPTER 1

HALOCARBON REFRIGERATION SYSTEMS Application................................................................................. 1.1 System Safety.............................................................................. 1.2 Basic Piping Principles ............................................................. 1.2 Refrigerant Line Sizing .............................................................. 1.3 Piping at Multiple Compressors .............................................. 1.20 Piping at Various System Components .................................... 1.21 Discharge (Hot-Gas) Lines ...................................................... 1.24 Defrost Gas Supply Lines......................................................... 1.26

Heat Exchangers and Vessels ................................................... Refrigeration Accessories ........................................................ Pressure Control for Refrigerant Condensers.......................... Keeping Liquid from Crankcase During Off Cycles ................ Hot-Gas Bypass Arrangements ................................................ Minimizing Refrigerant Charge in Commercial Systems ......... Refrigerant Retrofitting ............................................................ Temperature Glide....................................................................

R

chlorine could cause to the ozone layer in the stratosphere. This publication eventually led to the Montreal Protocol Agreement in 1987 and its subsequent revisions, which restricted the production and use of chlorinated halocarbon (CFC and HCFC) refrigerants. All CFC refrigerant production was phased out in the United States at the beginning of 1996. The development of replacement HFC, thirdgeneration refrigerants ensued following these restrictions (Calm 2008). Although HFC refrigerants do not contain chlorine and thus have no effect on stratospheric ozone, they have come under heavy scrutiny because of their global warming potential (GWP): like CFCs and HFCs, they are greenhouse gases, and can trap radiant energy (IPPC 1990). HFO refrigerants, however, have significantly lower GWP values, and are being developed and promoted as alternatives to HFC refrigerants. A successful refrigeration system depends on good piping design and an understanding of the required accessories. This chapter covers the fundamentals of piping and accessories in halocarbon refrigerant systems. Hydrocarbon refrigerant pipe friction data can be found in petroleum industry handbooks. Use the refrigerant properties and information in Chapters 3, 29, and 30 of the 2013 ASHRAE Handbook—Fundamentals to calculate friction losses. For information on refrigeration load, see Chapter 24. For R-502 information, refer to the 1998 ASHRAE Handbook—Refrigeration.

EFRIGERATION is the process of moving heat from one location to another by use of refrigerant in a closed cycle. Oil management; gas and liquid separation; subcooling, superheating, desuperheating, and piping of refrigerant liquid, gas, and two-phase flow are all part of refrigeration. Applications include air conditioning, commercial refrigeration, and industrial refrigeration. This chapter focuses on systems that use halocarbons (halogenated hydrocarbons) as refrigerants. The most commonly used halogen refrigerants are chlorine (Cl) and fluorine (F). Halocarbon refrigerants are classified into four groups: chlorofluorocarbons (CFCs), which contain carbon, chlorine, and fluorine; hydrochlorofluorocarbons (HCFCs), which consist of carbon, hydrogen, chlorine, and fluorine; hydrofluorocarbons (HFCs), which contain carbon, hydrogen, and fluorine; and hydrofluoroolefins (HFOs), which are HFC refrigerants derived from an alkene (olefin; i.e., an unsaturated compound having at least one carbon-to-carbon double bond). Examples of these refrigerants can be found in Chapter 29 of the 2013 ASHRAE Handbook—Fundamentals. Desired characteristics of a halocarbon refrigeration system may include • Year-round operation, regardless of outdoor ambient conditions • Possible wide load variations (0 to 100% capacity) during short periods without serious disruption of the required temperature levels • Frost control for continuous-performance applications • Oil management for different refrigerants under varying load and temperature conditions • A wide choice of heat exchange methods (e.g., dry expansion, liquid overfeed, or flooded feed of the refrigerants) and use of secondary coolants such as salt brine, alcohol, glycol, and carbon dioxide. • System efficiency, maintainability, and operating simplicity • Operating pressures and pressure ratios that might require multistaging, cascading, and so forth Development of halocarbon refrigerants dates back to the 1920s. The main refrigerants used then were ammonia (R-717), chloromethane (R-40), and sulfur dioxide (R-764), all of which have some degree of toxicity and/or flammability. These first-generation refrigerants were an impediment to Frigidaire’s plans to expand into refrigeration and air conditioning, so Frigidaire and DuPont collaborated to develop safer refrigerants. In 1928, Thomas Midgley, Jr., of Frigidaire and his colleagues developed the first commercially available CFC refrigerant, dichlorodifluoromethane (R-12) (Giunta 2006). Chlorinated halocarbon refrigerants represent the second generation of refrigerants (Calm 2008). Concern about the use of halocarbon refrigerants began with a 1974 paper by two University of California professors, Frank Rowland and Mario Molina, in which they highlighted the damage

The preparation of this chapter is assigned to TC 10.3, Refrigerant Piping.

1.1

1.26 1.29 1.33 1.34 1.35 1.36 1.37 1.37

APPLICATION Beyond the operational system characteristics described previously, political and environmental factors may need to be accounted for when designing, building, and installing halocarbon refrigeration systems. Heightened awareness of the impact halocarbon refrigerants have on ozone depletion and/or global warming has led to banning or phaseouts of certain refrigerants. Some end users are concerned about the future cost and availability of these refrigerants, and may fear future penalties that may come with owning and operating systems that use halocarbons. Therefore, many owners, engineers, and manufacturers seek to reduce charge and build tighter systems to reduce the total system charge on site and ensure that less refrigerant is released into the atmosphere. However, halocarbon refrigeration systems are still widely used. Although CFCs have been banned and HCFCs are being phased out because of their ODP, HFCs, which have a global warming potential (GWP), are still used in new installations and will continue to be used as the industries transition to natural or other refrigerants that may boast a reduced GWP. Table 1 in Chapter 3 lists commonly used refrigerants and their corresponding GWP values. Use of indirect and cascade systems to reduce the total amount of refrigerant has become increasingly popular. These systems also reduce the possibility for leakage because large amounts of interconnecting piping between the compressors and the heat load are

1.2

2014 ASHRAE Handbook—Refrigeration (SI) Table 1

Recommended Gas Line Velocities

Suction line Discharge line

4.5 to 20 m/s 10 to 18 m/s

Fig. 1 Flow Rate per Ton of Refrigeration for Refrigerant 22

replaced mainly with glycol or CO2 piping. (See Chapter 9 for more information on refrigerant containment, recovery, recycling, and reclamation.)

SYSTEM SAFETY ASHRAE Standard 15 and ASME Standard B31.5 should be used as guides for safe practice because they are the basis of most municipal and state codes. However, some ordinances require heavier piping and other features. The designer should know the specific requirements of the installation site. Only A106 Grade A or B or A53 Grade A or B should be considered for steel refrigerant piping. The rated internal working pressure for Type L copper tubing decreases with (1) increasing metal operating temperature, (2) increasing tubing size (OD), and (3) increasing temperature of joining method. Hot methods used to join drawn pipe (e.g., brazing, welding) produce joints as strong as surrounding pipe, but reduce the strength of the heated pipe material to that of annealed material. Particular attention should be paid when specifying copper in conjunction with newer, high-pressure refrigerants (e.g., R-404A, R-507A, R-410A, R-407C) because some of these refrigerants can achieve operating pressures as high as 3450 kPa and operating temperatures as high as 150°C at a typical saturated condensing condition of 55°C. Concentration calculations, based on the amount of refrigerant in the system and the volume of the space where it is installed, are needed to identify what safety features are required by the appropriate codes. Whenever allowable concentration limits of the refrigerant may be exceeded in occupied spaces, additional safety measures (e.g., leak detection, alarming, ventilation, automatic shut-off controls) are typically required. Note that, because halocarbon refrigerants are heavier than air, leak detection sensors should be placed at lower elevations in the space (typically 300 mm from the floor).

Fig. 1 Fig. 2 134a

Flow Rate per Kilowatt of Refrigeration for Refrigerant 22

Flow Rate per Ton of Refrigeration for Refrigerant

BASIC PIPING PRINCIPLES The design and operation of refrigerant piping systems should (1) ensure proper refrigerant feed to evaporators, (2) provide practical refrigerant line sizes without excessive pressure drop, (3) prevent excessive amounts of lubricating oil from being trapped in any part of the system, (4) protect the compressor at all times from loss of lubricating oil, (5) prevent liquid refrigerant or oil slugs from entering the compressor during operating and idle time, and (6) maintain a clean and dry system.

Refrigerant Line Velocities Economics, pressure drop, noise, and oil entrainment establish feasible design velocities in refrigerant lines (Table 1). Higher gas velocities are sometimes found in relatively short suction lines on comfort air-conditioning or other applications where the operating time is only 2000 to 4000 h per year and where low initial cost of the system may be more significant than low operating cost. Industrial or commercial refrigeration applications, where equipment runs almost continuously, should be designed with low refrigerant velocities for most efficient compressor performance and low equipment operating costs. An owning and operating cost analysis will reveal the best choice of line sizes. (See Chapter 37 of the 2011 ASHRAE Handbook—HVAC Applications for information on owning and operating costs.) Liquid lines from condensers to receivers should be sized for 0.5 m/s or less to ensure positive gravity flow without incurring back-up of liquid flow. Liquid lines from receiver to evaporator should be sized to maintain velocities below 1.5 m/s, thus minimizing or preventing liquid hammer when solenoids or other electrically operated valves are used.

Fig. 2

Flow Rate per Kilowatt of Refrigeration for Refrigerant 134a

Refrigerant Flow Rates Refrigerant flow rates for R-22 and R-134a are indicated in Figures 1 and 2. To obtain total system flow rate, select the proper rate value and multiply by system capacity. Enter curves using saturated refrigerant temperature at the evaporator outlet and actual liquid temperature entering the liquid feed device (including subcooling in condensers and liquid-suction interchanger, if used). Because Figures 1 and 2 are based on a saturated evaporator temperature, they may indicate slightly higher refrigerant flow rates than are actually in effect when suction vapor is superheated above the conditions mentioned. Refrigerant flow rates may be reduced approximately 0.5% for each 1 K increase in superheat in the evaporator. Suction-line superheating downstream of the evaporator from line heat gain from external sources should not be used to reduce evaluated mass flow, because it increases volumetric flow rate and line velocity per unit of evaporator capacity, but not mass flow rate. It should be considered when evaluating suction-line size for satisfactory oil return up risers. Suction gas superheating from use of a liquid-suction heat exchanger has an effect on oil return similar to that of suction-line superheating. The liquid cooling that results from the heat exchange

Halocarbon Refrigeration Systems

1.3

Table 2 Approximate Effect of Gas Line Pressure Drops on R-22 Compressor Capacity and Powera Capacity, %

Energy, %b

Suction Line 0 1 2

100 96.8 93.6

100 104.3 107.3

Discharge Line 0 1 2

100 99.2 98.4

100 102.7 105.7

Line Loss, K

aFor system operating at 5°C saturated evaporator temperature and 40°C saturated con-

densing temperature. percentage rated at kW (power)/kW (cooling).

bEnergy

reduces mass flow rate per unit of refrigeration. This can be seen in Figures 1 and 2 because the reduced temperature of the liquid supplied to the evaporator feed valve has been taken into account. Superheat caused by heat in a space not intended to be cooled is always detrimental because the volumetric flow rate increases with no compensating gain in refrigerating effect.

REFRIGERANT LINE SIZING In sizing refrigerant lines, cost considerations favor minimizing line sizes. However, suction and discharge line pressure drops cause loss of compressor capacity and increased power usage. Excessive liquid-line pressure drops can cause liquid refrigerant to flash, resulting in faulty expansion valve operation. Refrigeration systems are designed so that friction pressure losses do not exceed a pressure differential equivalent to a corresponding change in the saturation boiling temperature. The primary measure for determining pressure drops is a given change in saturation temperature.

Pressure Drop Considerations Pressure drop in refrigerant lines reduces system efficiency. Correct sizing must be based on minimizing cost and maximizing efficiency. Table 2 shows the approximate effect of refrigerant pressure drop on an R-22 system operating at a 5°C saturated evaporator temperature with a 40°C saturated condensing temperature. Pressure drop calculations are determined as normal pressure loss associated with a change in saturation temperature of the refrigerant. Typically, the refrigeration system is sized for pressure losses of 1 K or less for each segment of the discharge, suction, and liquid lines. Liquid Lines. Pressure drop should not be so large as to cause gas formation in the liquid line, insufficient liquid pressure at the liquid feed device, or both. Systems are normally designed so that pressure drop in the liquid line from friction is not greater than that corresponding to about a 0.5 to 1 K change in saturation temperature. See Tables 3 to 9 for liquid-line sizing information. Liquid subcooling is the only method of overcoming liquid line pressure loss to guarantee liquid at the expansion device in the evaporator. If subcooling is insufficient, flashing occurs in the liquid line and degrades system efficiency. Friction pressure drops in the liquid line are caused by accessories such as solenoid valves, filter-driers, and hand valves, as well as by the actual pipe and fittings between the receiver outlet and the refrigerant feed device at the evaporator. Liquid-line risers are a source of pressure loss and add to the total loss of the liquid line. Loss caused by risers is approximately 11.3 kPa per metre of liquid lift. Total loss is the sum of all friction losses plus pressure loss from liquid risers. Example 1 illustrates the process of determining liquid-line size and checking for total subcooling required. Example 1. An R-22 refrigeration system using copper pipe operates at 5°C evaporator and 40°C condensing. Capacity is 14 kW, and the liquid

line is 50 m equivalent length with a riser of 6 m. Determine the liquidline size and total required subcooling. Solution: From Table 3, the size of the liquid line at 1 K drop is 15 mm OD. Use the equation in Note 3 of Table 3 to compute actual temperature drop. At 14 kW, Actual temperature drop = (50  0.02)(14.0/21.54)1.8 Estimated friction loss = 0.46(50 × 0.749) Loss for the riser = 6  11.3 Total pressure losses = 67.8 + 17.2 Saturation pressure at 40°C condensing (see R-22 properties in Chapter 30, 2013 ASHRAE Handbook—Fundamentals) Initial pressure at beginning of liquid line Total liquid line losses Net pressure at expansion device The saturation temperature at 1448.6 kPa is 37.7°C. Required subcooling to overcome the liquid losses

= 0.46 K = 17.2 kPa = 67.8 kPa = 85.0 kPa = 1533.6 kPa

1533.6 kPa – 85.0 kPa = 1448.6 kPa = (40.0 – 37.7) or 2.3 K

Refrigeration systems that have no liquid risers and have the evaporator below the condenser/receiver benefit from a gain in pressure caused by liquid weight and can tolerate larger friction losses without flashing. Regardless of the liquid-line routing when flashing occurs, overall efficiency is reduced, and the system may malfunction. The velocity of liquid leaving a partially filled vessel (e.g., receiver, shell-and-tube condenser) is limited by the height of the liquid above the point at which the liquid line leaves the vessel, whether or not the liquid at the surface is subcooled. Because liquid in the vessel has a very low (or zero) velocity, the velocity V in the liquid line (usually at the vena contracta) is V 2 = 2gh, where h is the liquid height in the vessel. Gas pressure does not add to the velocity unless gas is flowing in the same direction. As a result, both gas and liquid flow through the line, limiting the rate of liquid flow. If this factor is not considered, excess operating charges in receivers and flooding of shell-and-tube condensers may result. No specific data are available to precisely size a line leaving a vessel. If the height of liquid above the vena contracta produces the desired velocity, liquid leaves the vessel at the expected rate. Thus, if the level in the vessel falls to one pipe diameter above the bottom of the vessel from which the liquid line leaves, the capacity of copper lines for R-22 at 6.4 g/s per kilowatt of refrigeration is approximately as follows: OD, mm 28 35 42 54 67 79 105

kW 49 88 140 280 460 690 1440

The whole liquid line need not be as large as the leaving connection. After the vena contracta, the velocity is about 40% less. If the line continues down from the receiver, the value of h increases. For a 700 kW capacity with R-22, the line from the bottom of the receiver should be about 79 mm. After a drop of 1300 mm, a reduction to 54 mm is satisfactory. Suction Lines. Suction lines are more critical than liquid and discharge lines from a design and construction standpoint. Refrigerant lines should be sized to (1) provide a minimum pressure drop at full load, (2) return oil from the evaporator to the compressor under minimum load conditions, and (3) prevent oil from draining from an active evaporator into an idle one. A pressure drop in the suction line reduces a system’s capacity because it forces the compressor to operate at a lower suction pressure to maintain a desired evaporating temperature in the coil. The suction line is normally

1.4

2014 ASHRAE Handbook—Refrigeration (SI)

Table 3 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 22 (Single- or High-Stage Applications)

Nominal Line OD, mm

–40 196

12 15 18 22 28 35 42 54 67 79 105

0.32 0.61 1.06 1.88 3.73 6.87 11.44 22.81 40.81 63.34 136.0

10 15 20 25 32 40 50 65 80 100

0.47 0.88 1.86 3.52 7.31 10.98 21.21 33.84 59.88 122.3

Suction Lines (t = 0.04 K/m) Discharge Lines Saturated Suction Temperature, °C (t = 0.02 K/m, p = 74.90) –30 –20 –5 5 Saturated Suction Corresponding p, Pa/m Temperature, °C 277 378 572 731 –40 –20 5 TYPE L COPPER LINE 0.50 0.75 1.28 1.76 2.30 2.44 2.60 0.95 1.43 2.45 3.37 4.37 4.65 4.95 1.66 2.49 4.26 5.85 7.59 8.06 8.59 2.93 4.39 7.51 10.31 13.32 14.15 15.07 5.82 8.71 14.83 20.34 26.24 27.89 29.70 10.70 15.99 27.22 37.31 48.03 51.05 54.37 17.80 26.56 45.17 61.84 79.50 84.52 90.00 35.49 52.81 89.69 122.7 157.3 167.2 178.1 63.34 94.08 159.5 218.3 279.4 297.0 316.3 98.13 145.9 247.2 337.9 431.3 458.5 488.2 210.3 312.2 527.8 721.9 919.7 977.6 1041.0 STEEL LINE 0.72 1.06 1.78 2.42 3.04 3.23 3.44 1.35 1.98 3.30 4.48 5.62 5.97 6.36 2.84 4.17 6.95 9.44 11.80 12.55 13.36 5.37 7.87 13.11 17.82 22.29 23.70 25.24 11.12 16.27 27.11 36.79 46.04 48.94 52.11 16.71 24.45 40.67 55.21 68.96 73.31 78.07 32.23 47.19 78.51 106.4 132.9 141.3 150.5 51.44 75.19 124.8 169.5 211.4 224.7 239.3 90.95 132.8 220.8 299.5 373.6 397.1 422.9 185.6 270.7 450.1 610.6 761.7 809.7 862.2

Notes: 1. Table capacities are in kilowatts of refrigeration. p = pressure drop per unit equivalent length of line, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le

Liquid Lines See note a Velocity = 0.5 m/s

t = 0.02 K/m p = 749

7.08 11.49 17.41 26.66 44.57 70.52 103.4 174.1 269.9 376.5 672.0

11.24 21.54 37.49 66.18 131.0 240.7 399.3 794.2 1415.0 2190.9 4697.0

10.66 16.98 29.79 48.19 83.56 113.7 187.5 267.3 412.7 711.2

15.96 29.62 62.55 118.2 244.4 366.6 707.5 1127.3 1991.3 4063.2

4. Values based on 40°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures.

Table L Actual t 0.55 Line capacity = Table capacity  ----------------------e-  -----------------------   Actual L e Table t  3. Saturation temperature t for other capacities and equivalent lengths Le Actual L Actual capacity 1.8 t = Table t  -----------------------e  -------------------------------------   Table L e   Table capacity  a Sizing is recommended where any gas generated in receiver must return up condensate line to condenser without restricting condensate flow. Water-cooled condensers, where receiver ambient temperature may be higher than refrigerant condensing temperature, fall into this category.

Condensing Temperature, °C 20 30 40 50

Suction Line 1.18 1.10 1.00 0.91

Discharge Line 0.80 0.88 1.00 1.11

pressure drop p is conservative; if subcooling is substantial or line is short, a smaller size line may be used. Applications with very little subcooling or very long lines may require a larger line.

b Line

Table 4 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 22 (Intermediate- or Low-Stage Duty)

Nominal Type L Copper Line OD, mm 12 15 18 22 28 35 42 54 67 79 105 130 156

–70 31.0 0.09 0.17 0.29 0.52 1.05 1.94 3.26 6.54 11.77 18.32 39.60 70.87 115.74

Suction Lines (t = 0.04 K/m) Saturated Suction Temperature, °C –60 –50 –40 Corresponding p, Pa/m 51.3 81.5 121 0.16 0.27 0.47 0.31 0.52 0.90 0.55 0.91 1.57 0.97 1.62 2.78 1.94 3.22 5.52 3.60 5.95 10.17 6.00 9.92 16.93 12.03 19.83 33.75 21.57 35.47 60.38 33.54 55.20 93.72 72.33 118.66 201.20 129.17 211.70 358.52 210.83 344.99 583.16

Notes: 1. Table capacities are in kilowatts of refrigeration. p = pressure drop per equivalent line length, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le Table L e Actual t 0.55 Line capacity = Table capacity  -----------------------  -----------------------  Actual L e Table t 3. Saturation temperature t for other capacities and equivalent lengths Le Actual L Actual capacity 1.8 t = Table t  -----------------------e   -------------------------------------  Table L e Table capacity *See the section on Pressure Drop Considerations.

–30 228 0.73 1.39 2.43 4.30 8.52 15.68 26.07 51.98 92.76 143.69 308.02 548.66 891.71

Discharge Lines* 0.74 1.43 2.49 4.41 8.74 16.08 26.73 53.28 95.06 174.22 316.13 561.89 915.02

Liquid Lines

See Table 3

4. Refer to refrigerant property tables (Chapter 30 of the 2013 ASHRAE Handbook—Fundamentals) for pressure drop corresponding to t. 5. Values based on –15°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures. Condensing Temperature, °C Suction Line Discharge Line –30 1.08 0.74 –20 1.03 0.91 –10 0.98 1.09 0 0.91 1.29

Halocarbon Refrigeration Systems

1.5

Table 5 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 134a (Single- or High-Stage Applications) Suction Lines (t = 0.04 K/m)

Nominal Line OD, mm

–10 318

Liquid Lines

Discharge Lines (t = 0.02 K/m, p = 538 Pa/m)

Saturated Suction Temperature, °C –5 0 5 Corresponding p, Pa/m 368 425 487

10

Saturated Suction Temperature, °C –10 0 10

555

See note a Velocity = 0.5 m/s

t = 0.02 K/m p = 538 Pa/m

TYPE L COPPER LINE 12 15 18 22 28 35 42 54 67 79 105

0.62 1.18 2.06 3.64 7.19 13.20 21.90 43.60 77.70 120.00 257.00

0.76 1.45 2.52 4.45 8.80 16.10 26.80 53.20 94.60 147.00 313.00

0.92 1.76 3.60 5.40 10.70 19.50 32.40 64.40 115.00 177.00 379.00

1.11 2.12 3.69 6.50 12.80 23.50 39.00 77.30 138.00 213.00 454.00

10 15 20 25 32 40 50 65 80 100

0.87 1.62 3.41 6.45 13.30 20.00 38.60 61.50 109.00 222.00

1.06 1.96 4.13 7.81 16.10 24.20 46.70 74.30 131.00 268.00

1.27 2.36 4.97 9.37 19.40 29.10 56.00 89.30 158.00 322.00

1.52 2.81 5.93 11.20 23.10 34.60 66.80 106.00 288.00 383.00

1.33 2.54 4.42 7.77 15.30 28.10 46.50 92.20 164.00 253.00 541.00

1.69 3.23 5.61 9.87 19.50 35.60 59.00 117.00 208.00 321.00 686.00

1.77 3.37 5.85 10.30 20.30 37.20 61.60 122.00 217.00 335.00 715.00

1.84 3.51 6.09 10.70 21.10 38.70 64.10 127.00 226.00 349.00 744.00

6.51 10.60 16.00 24.50 41.00 64.90 95.20 160.00 248.00 346.00 618.00

8.50 16.30 28.40 50.10 99.50 183.00 304.00 605.00 1080.00 1670.00 3580.00

2.28 4.22 8.88 16.70 34.60 51.90 100.00 159.00 281.00 573.00

2.38 4.40 9.26 17.50 36.10 54.10 104.00 166.00 294.00 598.00

2.47 4.58 9.64 18.20 37.50 56.30 108.00 173.00 306.00 622.00

9.81 15.60 27.40 44.40 76.90 105.00 173.00 246.00 380.00 655.00

12.30 22.80 48.20 91.00 188.00 283.00 546.00 871.00 1540.00 3140.00

STEEL LINE 1.80 3.34 7.02 13.30 27.40 41.00 79.10 126.00 223.00 454.00

4. Values based on 40°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures.

Notes: 1. Table capacities are in kilowatts of refrigeration. p = pressure drop per equivalent line length, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le

Condensing Temperature, °C 20 30 40 50

Table L Actual t 0.55 Line capacity = Table capacity  ----------------------e-  -----------------------   Actual L e Table t  3. Saturation temperature t for other capacities and equivalent lengths Le Actual L Actual capacity 1.8 t = Table t  -----------------------e   -------------------------------------   Table L e   Table capacity  a Sizing

is recommended where any gas generated in receiver must return up condensate line to condenser without restricting condensate flow. Water-cooled condensers, where receiver ambient temperature may be higher than refrigerant condensing temperature, fall into this category.

sized to have a pressure drop from friction no greater than the equivalent of about a 1 K change in saturation temperature. See Tables 3 to 15 for suction line sizing information. At suction temperatures lower than 5°C, the pressure drop equivalent to a given temperature change decreases. For example, at –40°C suction with R-22, the pressure drop equivalent to a 1 K change in saturation temperature is about 4.9 kPa. Therefore, low-temperature lines must be sized for a very low pressure drop, or higher equivalent temperature losses, with resultant loss in equipment capacity, must be accepted. For very low pressure drops, any suction or hot-gas risers must be sized properly to ensure oil entrainment up the riser so that oil is always returned to the compressor. Where pipe size must be reduced to provide sufficient gas velocity to entrain oil up vertical risers at partial loads, greater pressure drops are imposed at full load. These can usually be compensated for by oversizing the horizontal and down run lines and components. Discharge Lines. Pressure loss in hot-gas lines increases the required compressor power per unit of refrigeration and decreases compressor capacity. Table 2 illustrates power losses for an R-22 system at 5°C evaporator and 40°C condensing temperature. Pressure drop is minimized by generously sizing lines for low friction losses, but still maintaining refrigerant line velocities to entrain and carry oil along at all loading conditions. Pressure drop is normally

Suction Line 1.239 1.120 1.0 0.888

Discharge Line 0.682 0.856 1.0 1.110

pressure drop p is conservative; if subcooling is substantial or line is short, a smaller size line may be used. Applications with very little subcooling or very long lines may require a larger line.

b Line

designed not to exceed the equivalent of a 1 K change in saturation temperature. Recommended sizing tables are based on a 0.02 K/m change in saturation temperature.

Location and Arrangement of Piping Refrigerant lines should be as short and direct as possible to minimize tubing and refrigerant requirements and pressure drops. Plan piping for a minimum number of joints using as few elbows and other fittings as possible, but provide sufficient flexibility to absorb compressor vibration and stresses caused by thermal expansion and contraction. Arrange refrigerant piping so that normal inspection and servicing of the compressor and other equipment is not hindered. Do not obstruct the view of the oil-level sight glass or run piping so that it interferes with removing compressor cylinder heads, end bells, access plates, or any internal parts. Suction-line piping to the compressor should be arranged so that it will not interfere with removal of the compressor for servicing. Provide adequate clearance between pipe and adjacent walls and hangers or between pipes for insulation installation. Use sleeves that are sized to allow installation of both pipe and insulation through floors, walls, or ceilings. Set these sleeves before pouring concrete or erecting brickwork.

1.6 Run piping so that it does not interfere with passages or obstruct headroom, windows, and doors. Refer to ASHRAE Standard 15 and other governing local codes for restrictions that may apply.

Protection Against Damage to Piping Protection against damage is necessary, particularly for small lines, which have a false appearance of strength. Where traffic is heavy, provide protection against impact from carelessly handled hand trucks, overhanging loads, ladders, and fork trucks.

Piping Insulation All piping joints and fittings should be thoroughly leak-tested before insulation is sealed. Suction lines should be insulated to prevent sweating and heat gain. Insulation covering lines on which moisture can condense or lines subjected to outdoor conditions must be vapor sealed to prevent any moisture travel through the insulation or condensation in the insulation. Many commercially available types are provided with an integral waterproof jacket for this purpose. Although the liquid line ordinarily does not require insulation, suction and liquid lines can be insulated as a unit on installations where the two lines are clamped together. When it passes through a warmer area, the liquid line should be insulated to minimize heat gain. Hotgas discharge lines usually are not insulated; however, they should be insulated if necessary to prevent injury from high-temperature surfaces, or if the heat dissipated is objectionable (e.g., in systems that use heat reclaim). In this case, discharge lines upstream of the heat reclaim heat exchanger should be insulated. Downstream lines (between the heat reclaim heat exchanger and condenser) do not need to be insulated unless necessary to prevent the refrigerant from condensing prematurely. Also, indoor hot-gas discharge line insulation does not need a tight vapor seal because moisture condensation is not an issue. All joints and fittings should be covered, but it is not advisable to do so until the system has been thoroughly leak-tested. See Chapter 10 for additional information.

Vibration and Noise in Piping Vibration transmitted through or generated in refrigerant piping and the resulting objectionable noise can be eliminated or minimized by proper piping design and support. Two undesirable effects of vibration of refrigerant piping are (1) physical damage to the piping, which can break brazed joints and, consequently, lose charge; and (2) transmission of noise through the piping itself and through building construction that may come into direct contact with the piping. In refrigeration applications, piping vibration can be caused by rigid connection of the refrigerant piping to a reciprocating compressor. Vibration effects are evident in all lines directly connected to the compressor or condensing unit. It is thus impossible to eliminate vibration in piping; it is only possible to mitigate its effects. Flexible metal hose is sometimes used to absorb vibration transmission along smaller pipe sizes. For maximum effectiveness, it should be installed parallel to the crankshaft. In some cases, two isolators may be required, one in the horizontal line and the other in the vertical line at the compressor. A rigid brace on the end of the flexible hose away from the compressor is required to prevent vibration of the hot-gas line beyond the hose. Flexible metal hose is not as efficient in absorbing vibration on larger pipes because it is not actually flexible unless the ratio of length to diameter is relatively great. In practice, the length is often limited, so flexibility is reduced in larger sizes. This problem is best solved by using flexible piping and isolation hangers where the piping is secured to the structure. When piping passes through walls, through floors, or inside furring, it must not touch any part of the building and must be supported only by the hangers (provided to avoid transmitting vibration

2014 ASHRAE Handbook—Refrigeration (SI) to the building); this eliminates the possibility of walls or ceilings acting as sounding boards or diaphragms. When piping is erected where access is difficult after installation, it should be supported by isolation hangers. Vibration and noise from a piping system can also be caused by gas pulsations from the compressor operation or from turbulence in the gas, which increases at high velocities. It is usually more apparent in the discharge line than in other parts of the system. When gas pulsations caused by the compressor create vibration and noise, they have a characteristic frequency that is a function of the number of gas discharges by the compressor on each revolution. This frequency is not necessarily equal to the number of cylinders, because on some compressors two pistons operate together. It is also varied by the angular displacement of the cylinders, such as in V-type compressors. Noise resulting from gas pulsations is usually objectionable only when the piping system amplifies the pulsation by resonance. On single-compressor systems, resonance can be reduced by changing the size or length of the resonating line or by installing a properly sized hot-gas muffler in the discharge line immediately after the compressor discharge valve. On a paralleled compressor system, a harmonic frequency from the different speeds of multiple compressors may be apparent. This noise can sometimes be reduced by installing mufflers. When noise is caused by turbulence and isolating the line is not effective enough, installing a larger-diameter pipe to reduce gas velocity is sometimes helpful. Also, changing to a line of heavier wall or from copper to steel to change the pipe natural frequency may help.

Refrigerant Line Capacity Tables Tables 3 to 9 show line capacities in kilowatts of refrigeration for R-22, R-134A, R-404A, R-507A, R-410A, and R-407C. Capacities in the tables are based on the refrigerant flow that develops a friction loss, per metre of equivalent pipe length, corresponding to a 0.04 K change in the saturation temperature (t) in the suction line, and a 0.02 K change in the discharge line. The capacities shown for liquid lines are for pressure losses corresponding to 0.02 and 0.05 K/m change in saturation temperature and also for velocity corresponding to 0.5 m/s. Tables 10 to 15 show capacities for the same refrigerants based on reduced suction line pressure loss corresponding to 0.02 and 0.01 K/m equivalent length of pipe. These tables may be used when designing system piping to minimize suction line pressure drop. The refrigerant line sizing capacity tables are based on the DarcyWeisbach relation and friction factors as computed by the Colebrook function (Colebrook 1938, 1939). Tubing roughness height is 1.5 m for copper and 46 m for steel pipe. Viscosity extrapolations and adjustments for pressures other than 101.325 kPa were based on correlation techniques as presented by Keating and Matula (1969). Discharge gas superheat was 45 K for R-134a and 60 K for R-22. The refrigerant cycle for determining capacity is based on saturated gas leaving the evaporator. The calculations neglect the presence of oil and assume nonpulsating flow. For additional charts and discussion of line sizing refer to Atwood (1990), Timm (1991), and Wile (1977).

Equivalent Lengths of Valves and Fittings Refrigerant line capacity tables are based on unit pressure drop per metre length of straight pipe, or per combination of straight pipe, fittings, and valves with friction drop equivalent to a metre of straight pipe. Generally, pressure drop through valves and fittings is determined by establishing the equivalent straight length of pipe of the same size with the same friction drop. Line sizing tables can then be used directly. Tables 16 to 18 give equivalent lengths of straight pipe for various fittings and valves, based on nominal pipe sizes. The following example illustrates the use of various tables and charts to size refrigerant lines.

Suction Lines (t = 0.04 K/m)

Discharge Lines (t = 0.02 K/m, p = 74.90)

165.5 0.16 0.30 0.53 0.94 1.86 3.43 5.71 11.37 20.31 31.54 67.66 120.40 195.94 401.89 715.93

Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 240.6 337.2 455.1 679.1 0.27 0.43 0.67 1.19 0.52 0.83 1.28 2.27 0.90 1.45 2.22 3.94 1.59 2.55 3.91 6.93 3.14 5.04 7.72 13.66 5.78 9.26 14.15 25.00 9.61 15.36 23.46 41.32 19.12 30.50 46.57 81.90 34.10 54.30 82.75 145.45 52.78 84.12 128.09 224.52 113.08 179.89 273.26 478.70 201.19 319.22 484.40 847.54 326.58 518.54 785.73 1372.94 669.47 1059.73 1607.24 2805.00 1189.91 1885.42 2851.68 4974.31

Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 875.6 875.6 875.6 875.6 1.87 2.00 2.13 2.31 3.55 3.81 4.05 4.40 6.16 6.59 7.02 7.62 10.79 11.56 12.30 13.36 21.23 22.74 24.21 26.29 38.78 41.54 44.23 48.03 64.15 68.72 73.16 79.45 126.86 135.89 144.67 157.11 225.07 241.08 256.66 278.73 346.97 371.66 395.67 429.70 738.92 791.51 842.65 915.11 1309.04 1402.20 1492.80 1621.17 2116.83 2267.48 2413.98 2621.57 4317.73 4625.02 4923.84 5347.26 7641.29 8185.11 8713.94 9463.30

0.16 0.31 0.70 1.37 2.95 4.49 10.47 16.68 29.51 60.26 108.75 176.25 360.41 652.69 1044.01 1351.59 1947.52

0.26 0.51 1.15 2.25 4.83 7.38 17.16 27.33 48.38 98.60 177.97 287.77 589.35 1065.97 1705.26 2207.80 3176.58

Line Size Type L Copper, OD, mm 12 15 18 22 28 35 42 54 67 79 105 130 156 206 257 Steel mm SCH 10 80 15 80 20 80 25 80 32 80 40 80 50 40 65 40 80 40 100 40 125 40 150 40 200 40 250 40 300 IDb 350 30 400 30 a Sizing

–50

0.40 0.80 1.80 3.53 7.57 11.55 26.81 42.72 75.47 153.84 277.71 449.08 918.60 1661.62 2658.28 3436.53 4959.92

5

–50

863.2 1.69 3.22 5.57 9.79 19.25 35.17 58.16 114.98 203.96 314.97 670.69 1188.02 1921.03 3917.77 6949.80

875.6 1.73 3.29 5.71 10.00 19.68 35.96 59.48 117.62 208.67 321.69 685.09 1213.68 1962.62 4003.19 7084.63

0.61 1.05 1.46 1.49 1.20 2.07 2.88 2.94 2.70 4.66 6.48 6.61 5.30 9.13 12.68 12.95 11.35 19.57 27.20 27.72 17.29 29.81 41.42 42.22 40.20 69.20 96.18 98.04 63.93 110.18 152.98 155.95 112.96 194.49 270.35 275.59 230.29 396.56 550.03 560.67 415.78 714.27 991.91 1012.44 671.57 1155.17 1604.32 1635.36 1373.79 2363.28 3277.89 3341.30 2485.16 4275.41 5930.04 6044.77 3970.05 6830.36 9488.03 9671.59 5140.20 8843.83 12 266.49 12 503.79 7407.49 12 725.25 17 677.86 18 019.86

shown is recommended where any gas generated in receiver must return up condensate line to condenser without restricting condensate flow. Water-cooled condensers, where receiver ambient temperature may be higher than refrigerant condensing temperature, fall into this category. b Pipe inside diameter is same as nominal pipe size.

1.61 3.17 7.13 13.97 29.90 45.54 105.75 168.20 297.25 604.72 1091.99 1763.85 3603.84 6519.73 10 431.52 13 486.26 19 435.74

1.72 3.39 7.64 14.96 32.03 48.78 113.27 180.17 318.40 647.76 1169.71 1889.38 3860.32 6983.73 11 173.92 14 446.06 20 818.96

1.99 3.92 8.84 17.30 37.03 56.40 130.96 208.31 368.13 748.91 1352.37 2184.43 4463.15 8074.30 12 918.83 16 701.95 24 070.04

See note a 5 875.6 2.42 4.61 7.99 14.01 27.57 50.37 83.32 164.76 292.29 450.60 959.63 1700.03 2749.09 5607.37 9923.61

2.09 4.12 9.27 18.14 38.83 59.14 137.33 218.44 386.03 785.34 1418.15 2290.69 4680.25 8467.06 13 547.24 17 514.38 25 240.87

t = 0.02 K/m Drop Velocity = 0.5 m/s p = 875.6 4.1 8.0 6.7 15.3 10.1 26.6 15.5 46.8 26.0 92.5 41.1 169.3 60.3 280.4 101.4 556.9 157.3 989.8 219.3 1529.9 391.5 3264.9 607.3 5788.8 879.6 9382.5 1522.1 19 177.4 2366.6 33 992.3

4.6 7.6 14.1 23.4 41.8 57.5 109.2 155.7 240.5 414.3 650.6 940.3 1628.2 2566.4 3680.9 4487.7 5944.7

7.2 14.3 32.1 63.0 134.9 205.7 477.6 761.1 1344.9 2735.7 4939.2 7988.0 16 342.0 29 521.7 47 161.0 61 061.2 87 994.9

t = 0.05 K/m Drop p = 2189.1 13.3 25.2 43.7 76.7 151.1 276.3 456.2 903.2 1601.8 2473.4 5265.6 9335.2 15 109.7 30 811.3 54 651.2

11.5 22.7 51.1 100.0 214.0 326.5 758.2 1205.9 2131.2 4335.6 7819.0 12 629.7 25 838.1 46 743.9 74 677.7 96 691.3 139 346.8

4. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and Cond. Sucsaturated evaporator outlet temperature. Liquid capacity (kW) based Temp., tion on –5°C evaporator temperature. °C Line 5. Thermophysical properties and viscosity data based on calculations 20 1.344 from NIST REFPROP program Version 6.01. 30 1.177 6. For brazed Type L copper tubing larger than 28 mm OD for discharge 40 1.000 or liquid service, see Safety Requirements section. 7. Values are based on 40°C condensing temperature. Multiply table 50 0.809 capacities by the following factors for other condensing temperatures.

Discharge Line 0.812 0.906 1.000 1.035

1.7

Notes: 1. Table capacities are in kilowatts of refrigeration. p = pressure drop per unit equivalent length of line, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le Table L Actual t 0.55 Line capacity = Table capacity  ----------------------e-  -----------------------   Actual L e Table t  3. Saturation temperature t for other capacities and equivalent lengths Le Actual L e Actual capacity 1.8 t = Table t  -----------------------   -------------------------------------   Table L e   Table capacity 

1.83 3.61 8.14 15.93 34.10 51.94 120.59 191.81 338.98 689.61 1245.28 2011.45 4109.73 7434.94 11 895.85 15 379.40 22 164.04

Liquid Lines (40°C)

Halocarbon Refrigeration Systems

Table 6 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 404A (Single- or High-Stage Applications)

1.8

Table 7 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 507A (Single- or High-Stage Applications) Line Size

a Sizing

–50 173.7 0.16 0.31 0.55 0.97 1.91 3.52 5.86 11.68 20.86 32.31 69.31 123.41 200.86 412.07 733.42

0.16 0.31 0.71 1.40 3.01 4.59 10.69 17.06 30.20 61.60 111.17 179.98 368.55 666.52 1067.53 1380.23 1991.54

Discharge Lines (t = 0.02 K/m, p = 74.90) Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 896.3 896.3 896.3 896.3 1.86 2.00 2.13 2.32 3.54 3.80 4.05 4.41 6.12 6.57 7.01 7.63 10.73 11.52 12.29 13.37 21.12 22.67 24.18 26.31 38.58 41.42 44.17 48.07 63.82 68.52 73.07 79.52 126.22 135.51 144.51 157.26 223.53 239.99 255.92 278.52 345.26 370.68 395.29 430.19 733.87 787.90 840.21 914.39 1300.07 1395.78 1488.45 1619.87 2104.68 2259.62 2409.65 2622.39 4288.18 4603.88 4909.55 5343.00 7598.35 8157.74 8699.37 9467.42

0.26 0.52 1.17 2.29 4.93 7.52 17.50 27.88 49.26 100.39 181.20 292.99 600.02 1085.29 1736.16 2247.80 3239.15

0.41 0.81 1.83 3.58 7.68 11.72 27.25 43.32 76.63 156.20 281.64 455.44 931.61 1685.18 2695.93 3485.20 5030.17

shown is recommended where any gas generated in receiver must return up condensate line to condenser without restricting condensate flow. Water-cooled condensers, where receiver ambient temperature may be higher than refrigerant condensing temperature, fall into this category. b Pipe inside diameter is same as nominal pipe size.

0.62 1.21 2.74 5.36 11.50 17.54 40.71 64.81 114.52 233.20 421.03 680.92 1393.04 2516.51 4020.13 5205.04 7500.91

1.06 2.09 4.71 9.23 19.76 30.09 69.87 111.37 196.37 400.40 721.18 1166.35 2386.16 4316.82 6896.51 8929.47 12 848.49

5

–50

882.5 1.70 3.24 5.61 9.85 19.38 35.40 58.55 115.76 205.36 317.17 675.47 1194.03 1935.01 3937.64 6984.91

896.3 1.72 3.27 5.66 9.93 19.53 35.68 59.03 116.74 206.75 319.34 678.77 1202.46 1946.66 3966.22 7027.87

1.47 2.90 6.52 12.77 27.33 41.63 96.67 153.76 271.72 552.81 998.16 1612.43 3294.46 5960.02 9535.99 12 328.49 17 767.21

1.48 2.91 6.55 12.83 27.47 41.83 97.14 154.51 273.05 555.50 1003.06 1620.28 3310.49 5989.03 9582.41 12 388.50 17 853.70

1.60 3.15 7.09 13.87 29.70 45.23 105.02 167.05 295.22 600.59 1084.49 1751.80 3579.22 6475.19 10 360.26 13 394.13 19 302.97

1.72 3.38 7.61 14.89 31.88 48.56 112.76 179.35 316.95 644.81 1164.33 1880.77 3842.72 6951.89 11 122.98 14 380.20 20 724.05

Notes: 1. Table capacities are in kilowatts of refrigeration. p = pressure drop per unit equivalent length of line, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le Table L Actual t 0.55 Line capacity = Table capacity  ----------------------e-  -----------------------   Actual L e Table t  3. Saturation temperature t for other capacities and equivalent lengths Le Actual L Actual capacity 1.8 t = Table t  -----------------------e   -------------------------------------   Table L e   Table capacity 

1.83 3.60 8.11 15.88 34.00 51.78 120.24 191.26 338.00 687.62 1241.63 2005.64 4097.86 7413.46 11 861.49 15 334.97 22 100.02

1.99 3.92 8.83 17.28 37.00 56.35 130.86 208.14 367.84 748.33 1351.25 2182.72 4459.65 8067.98 12 908.71 16 688.86 24 051.18

Liquid Lines (40°C) See note a 5 896.3 2.43 4.63 8.01 14.04 27.63 50.47 83.50 165.12 292.43 451.67 960.06 1700.76 2753.36 5609.84 9940.23

2.09 4.12 9.27 18.15 38.85 59.17 137.39 218.54 386.21 785.70 1418.74 2291.73 4682.37 8470.90 13 553.39 17 522.33 25 252.33

t = 0.02 K/m t = 0.05 K/m Drop Drop Velocity = p = 896.3 p = 2240.8 0.5 m/s 4.0 7.9 13.0 6.5 15.0 24.7 9.8 26.1 42.8 15.0 45.9 75.1 25.1 90.5 147.8 39.7 165.6 270.0 58.2 274.8 447.1 98.0 544.0 883.9 151.9 967.0 1567.7 211.9 1497.3 2420.9 378.2 3189.5 5154.4 586.7 5666.6 9129.4 849.9 9175.8 14 793.3 30 099.9 1470.7 18 734.6 2286.7 33 285.5 53 389.2

4.4 7.4 13.6 22.6 40.3 55.6 105.5 150.4 232.3 400.3 628.6 908.5 1573.2 2479.7 3556.5 4336.1 5743.9

7.1 13.9 31.4 61.6 132.0 201.0 466.6 743.5 1313.9 2675.6 4825.1 7803.5 15 964.7 28 840.0 46 140.3 59 651.3 85 963.1

11.3 22.2 49.9 97.7 209.4 319.0 740.7 1178.1 2082.0 4235.5 7638.5 12 338.1 25 241.5 45 664.6 72 953.4 94 458.7 136 129.3

4. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and Cond. Sucsaturated evaporator outlet temperature. Liquid capacity (kW) based Temp., tion on –5°C evaporator temperature. °C Line 5. Thermophysical properties and viscosity data based on calculations 20 1.357 from NIST REFPROP program Version 6.01. 30 1.184 6. For brazed Type L copper tubing larger than 28 mm OD for discharge 40 1.000 or liquid service, see Safety Requirements section. 7. Values are based on 40°C condensing temperature. Multiply table 50 0.801 capacities by the following factors for other condensing temperatures.

Discharge Line 0.765 0.908 1.000 1.021

2014 ASHRAE Handbook—Refrigeration (SI)

Type L Copper, OD, mm 12 15 18 22 28 35 42 54 67 79 105 130 156 206 257 Steel mm SCH 10 80 15 80 20 80 25 80 32 80 40 80 50 40 65 40 80 40 100 40 125 40 150 40 200 40 250 40 300 IDb 350 30 400 30

Suction Lines (t = 0.04 K/m) Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 251.7 350.3 471.6 700.5 0.28 0.44 0.68 1.21 0.53 0.85 1.30 2.31 0.92 1.47 2.26 4.00 1.63 2.60 3.98 7.02 3.22 5.14 7.85 13.83 5.91 9.42 14.37 25.28 9.82 15.65 23.83 41.86 19.55 31.07 47.24 82.83 34.83 55.25 84.08 147.12 54.01 85.61 129.94 227.12 115.54 182.78 277.24 484.29 205.61 325.01 492.45 857.55 333.77 526.96 797.36 1389.26 683.01 1078.30 1631.18 2832.25 1216.78 1916.48 2891.11 5022.65

Line Size Type L Copper, OD, mm 12 15 18 22 28 35 42 54 67 79 105 130 156 206 257 Steel mm SCH 10 80 15 80 20 80 25 80 32 80 40 80 50 40 65 40 80 40 100 40 125 40 150 40 200 40 250 40 300 IDb 350 30 400 30 a Sizing

Discharge Lines (t = 0.02 K/m, p = 74.90)

Suction Lines (t = 0.04 K/m)

218.6 0.32 0.61 1.06 1.87 3.72 6.84 11.39 22.70 40.48 62.89 134.69 240.18 390.21 800.39 1427.49

Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 317.2 443.3 599.1 894.2 0.52 0.80 1.20 2.05 0.99 1.54 2.29 3.90 1.72 2.68 3.98 6.76 3.04 4.72 7.00 11.89 6.03 9.32 13.82 23.43 11.07 17.11 25.33 42.82 18.39 28.38 42.00 70.89 36.61 56.35 83.26 140.29 65.21 100.35 147.94 249.16 101.10 155.22 229.02 384.65 216.27 331.96 488.64 820.20 384.82 590.29 866.21 1452.34 625.92 957.07 1405.29 2352.81 1280.57 1956.28 2868.65 4796.70 2276.75 3480.75 5095.42 8506.22

1137.6 2.83 5.37 9.30 16.32 32.11 58.75 97.02 191.84 340.33 525.59 1119.32 1978.69 3206.57 6532.82 11 575.35

0.31 0.61 1.39 2.72 5.86 8.94 20.81 33.22 58.79 119.78 216.38 350.32 717.23 1297.30 2075.09 2686.45 3870.92

0.49 0.97 2.19 4.30 9.24 14.09 32.75 52.18 92.36 188.24 339.76 549.37 1125.10 2035.01 3255.45 4214.83 6064.31

2.44 4.80 10.81 21.16 45.30 68.99 160.19 254.80 450.29 916.08 1654.16 2672.01 5459.36 9876.55 15 802.42 20 429.97 29 442.67

–50

0.74 1.47 3.32 6.50 13.95 21.28 49.39 78.69 139.17 283.69 511.52 827.18 1692.00 3060.66 4896.39 6329.87 9135.88

1.08 2.14 4.82 9.45 20.26 30.91 71.75 114.11 201.84 411.01 742.06 1200.12 2451.89 4435.35 7085.49 9173.88 13 220.36

shown is recommended where any gas generated in receiver must return up condensate line to condenser without restricting condensate flow. Water-cooled condensers, where receiver ambient temperature may be higher than refrigerant condensing temperature, fall into this category. b Pipe inside diameter is same as nominal pipe size.

1.80 3.54 7.98 15.63 33.47 50.97 118.34 188.61 332.58 678.11 1221.40 1975.34 4041.21 7310.97 11 679.95 15 122.98 21 760.24

5

Liquid Lines (40°C)

1172.1 3.47 6.60 11.43 20.04 39.44 72.05 119.01 235.35 417.58 643.78 1371.21 2424.14 3928.86 7995.81 14 185.59

Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 1172.1 1172.1 1172.1 1172.1 3.60 3.73 3.84 4.00 6.85 7.09 7.31 7.60 11.87 12.29 12.67 13.16 20.81 21.54 22.20 23.08 40.95 42.39 43.70 45.42 74.82 77.46 79.84 82.98 123.57 127.93 131.87 137.06 244.38 253.00 260.80 271.06 433.60 448.89 462.73 480.93 668.47 692.05 713.37 741.44 1423.81 1474.02 1519.45 1579.22 2517.13 2605.89 2686.20 2791.88 4079.57 4223.44 4353.60 4524.87 8302.53 8595.32 8860.22 9208.77 14 729.76 15 249.20 15 719.17 16 337.55

1172.1 4.07 7.75 13.42 23.53 46.31 84.62 139.76 276.39 490.40 756.03 1610.30 2846.83 4613.92 9390.02 16 659.10

Velo- t = 0.02 K/m Drop city = 0.5 m/s p = 1179 6.2 14.3 10.1 27.2 15.4 47.3 23.5 83.0 39.3 163.7 62.2 299.6 91.3 495.7 153.7 982.0 238.2 1746.4 332.2 2695.2 592.9 5744.4 919.8 10 188.7 1332.3 16 502.3 2305.4 33 708.0 3584.6 59 763.6

2.98 5.87 13.21 25.86 55.37 84.33 195.83 311.49 550.47 1121.21 2022.16 3266.45 6673.89 12 073.76 19 317.94 24 974.96 35 992.70

3.10 6.09 13.72 26.85 57.50 87.57 203.34 323.43 571.59 1164.22 2099.73 3391.75 6929.91 12 536.92 20 059.00 25 933.02 37 373.41

3.50 6.89 15.52 30.37 65.03 99.04 229.98 365.80 646.46 1316.72 2374.75 3836.01 7837.60 14 179.04 22 686.37 29 329.79 42 268.67

6.9 11.5 21.3 35.5 63.2 87.1 165.4 235.8 364.2 627.6 985.4 1424.2 2466.2 3887.3 5575.3 6797.4 9004.3

–50

3.21 6.31 14.20 27.80 59.53 90.66 210.51 334.84 591.74 1205.28 2173.77 3511.36 7174.29 12 979.03 20 766.37 26 847.53 38 691.36

3.31 6.50 14.64 28.66 61.36 93.45 217.00 345.16 609.98 1242.42 2240.77 3619.58 7395.39 13 379.03 21 406.37 27 674.95 39 883.80

3.44 6.76 15.22 29.79 63.77 97.13 225.54 358.74 633.98 1291.30 2328.92 3761.97 7686.32 13 905.35 22 248.47 28 763.66 41 452.79

See note a 5

12.7 25.0 56.2 110.2 235.9 359.8 835.4 1328.6 2347.8 4787.0 8622.2 13 944.5 28 528.0 51 535.6 82 451.9 106 757.2 153 611.4

t = 0.05 K/m Drop p = 2935.8 23.5 44.6 77.2 135.3 266.4 486.0 804.1 1590.3 2816.7 4350.8 9249.0 16 386.3 26 500.6 53 996.3 95 683.0

Halocarbon Refrigeration Systems

Table 8 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 410A (Single- or High-Stage Applications)

20.1 39.6 89.1 174.5 568.9 1320.9 2101.0 3713.1 7562.8 13 639.9 22 032.9 45 016.9 81 440.3 130 304.0 168 461.9 242 779.1 Discharge Line 0.657 0.866 1.000 1.117

1.9

Notes: 4. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and Cond. Suc1. Table capacities are in kilowatts of refrigeration. saturated evaporator outlet temperature. Liquid capacity (kW) based Temp., tion p = pressure drop per unit equivalent length of line, Pa/m on –5°C evaporator temperature. °C Line t = corresponding change in saturation temperature, K/m 5. Thermophysical properties and viscosity data based on calculations 20 1.238 2. Line capacity for other saturation temperatures t and equivalent lengths Le from NIST REFPROP program Version 6.01. 30 1.122 6. For brazed Type L copper tubing larger than 15 mm OD for discharge Table L e Actual t 0.55 Line capacity = Table capacity  -----------------------  -----------------------  40 1.000 or liquid service, see Safety Requirements section. Actual L e Table t 7. Values are based on 40°C condensing temperature. Multiply table 50 0.867 3. Saturation temperature t for other capacities and equivalent lengths Le capacities by the following factors for other condensing temperatures. Actual L e   Actual capacity 1.8  t = Table t ---------------------------------------------------------- Table L e   Table capacity 

1.10

Table 9 Suction, Discharge, and Liquid Line Capacities in Kilowatts for Refrigerant 407C (Single- or High-Stage Applications) Line Size

a Sizing

–50 173.7 0.16 0.31 0.55 0.97 1.91 3.52 5.86 11.68 20.86 32.31 69.31 123.41 200.86 412.07 733.42

0.16 0.31 0.71 1.40 3.01 4.59 10.69 17.06 30.20 61.60 111.17 179.98 368.55 666.52 1067.53 1380.23 1991.54

Discharge Lines (t = 0.02 K/m, p = 74.90) Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 896.3 896.3 896.3 896.3 1.86 2.00 2.13 2.32 3.54 3.80 4.05 4.41 6.12 6.57 7.01 7.63 10.73 11.52 12.29 13.37 21.12 22.67 24.18 26.31 38.58 41.42 44.17 48.07 63.82 68.52 73.07 79.52 126.22 135.51 144.51 157.26 223.53 239.99 255.92 278.52 345.26 370.68 395.29 430.19 733.87 787.90 840.21 914.39 1300.07 1395.78 1488.45 1619.87 2104.68 2259.62 2409.65 2622.39 4288.18 4603.88 4909.55 5343.00 7598.35 8157.74 8699.37 9467.42

0.26 0.52 1.17 2.29 4.93 7.52 17.50 27.88 49.26 100.39 181.20 292.99 600.02 1085.29 1736.16 2247.80 3239.15

0.41 0.81 1.83 3.58 7.68 11.72 27.25 43.32 76.63 156.20 281.64 455.44 931.61 1685.18 2695.93 3485.20 5030.17

shown is recommended where any gas generated in receiver must return up condensate line to condenser without restricting condensate flow. Water-cooled condensers, where receiver ambient temperature may be higher than refrigerant condensing temperature, fall into this category. b Pipe inside diameter is same as nominal pipe size.

0.62 1.21 2.74 5.36 11.50 17.54 40.71 64.81 114.52 233.20 421.03 680.92 1393.04 2516.51 4020.13 5205.04 7500.91

1.06 2.09 4.71 9.23 19.76 30.09 69.87 111.37 196.37 400.40 721.18 1166.35 2386.16 4316.82 6896.51 8929.47 12 848.49

5

–50

882.5 1.70 3.24 5.61 9.85 19.38 35.40 58.55 115.76 205.36 317.17 675.47 1194.03 1935.01 3937.64 6984.91

896.3 1.72 3.27 5.66 9.93 19.53 35.68 59.03 116.74 206.75 319.34 678.77 1202.46 1946.66 3966.22 7027.87

1.47 2.90 6.52 12.77 27.33 41.63 96.67 153.76 271.72 552.81 998.16 1612.43 3294.46 5960.02 9535.99 12 328.49 17 767.21

1.48 2.91 6.55 12.83 27.47 41.83 97.14 154.51 273.05 555.50 1003.06 1620.28 3310.49 5989.03 9582.41 12 388.50 17 853.70

1.60 3.15 7.09 13.87 29.70 45.23 105.02 167.05 295.22 600.59 1084.49 1751.80 3579.22 6475.19 10 360.26 13 394.13 19 302.97

1.72 3.38 7.61 14.89 31.88 48.56 112.76 179.35 316.95 644.81 1164.33 1880.77 3842.72 6951.89 11 122.98 14 380.20 20 724.05

Notes: 1. Table capacities are in kilowatts of refrigeration. p = pressure drop per unit equivalent length of line, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le Table L Actual t 0.55 Line capacity = Table capacity  ----------------------e-  -----------------------   Actual L e Table t  3. Saturation temperature t for other capacities and equivalent lengths Le Actual L Actual capacity 1.8 t = Table t  -----------------------e   -------------------------------------   Table L e   Table capacity 

1.83 3.60 8.11 15.88 34.00 51.78 120.24 191.26 338.00 687.62 1241.63 2005.64 4097.86 7413.46 11 861.49 15 334.97 22 100.02

1.99 3.92 8.83 17.28 37.00 56.35 130.86 208.14 367.84 748.33 1351.25 2182.72 4459.65 8067.98 12 908.71 16 688.86 24 051.18

Liquid Lines (40°C) See note a 5 896.3 2.43 4.63 8.01 14.04 27.63 50.47 83.50 165.12 292.43 451.67 960.06 1700.76 2753.36 5609.84 9940.23

2.09 4.12 9.27 18.15 38.85 59.17 137.39 218.54 386.21 785.70 1418.74 2291.73 4682.37 8470.90 13 553.39 17 522.33 25 252.33

Velocity = 0.5 m/s 4.0 6.5 9.8 15.0 25.1 39.7 58.2 98.0 151.9 211.9 378.2 586.7 849.9 1470.7 2286.7

4.4 7.4 13.6 22.6 40.3 55.6 105.5 150.4 232.3 400.3 628.6 908.5 1573.2 2479.7 3556.5 4336.1 5743.9

t = 0.02 K/m Drop p = 896.3 7.9 15.0 26.1 45.9 90.5 165.6 274.8 544.0 967.0 1497.3 3189.5 5666.6 9175.8 18 734.6 33 285.5

7.1 13.9 31.4 61.6 132.0 201.0 466.6 743.5 1313.9 2675.6 4825.1 7803.5 15 964.7 28 840.0 46 140.3 59 651.3 85 963.1

t = 0.05 K/m Drop p = 2240.8 13.0 24.7 42.8 75.1 147.8 270.0 447.1 883.9 1567.7 2420.9 5154.4 9129.4 14 793.3 30 099.9 53 389.2

11.3 22.2 49.9 97.7 209.4 319.0 740.7 1178.1 2082.0 4235.5 7638.5 12 338.1 25 241.5 45 664.6 72 953.4 94 458.7 136 129.3

4. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and Cond. Sucsaturated evaporator outlet temperature. Liquid capacity (kW) based Temp., tion on –5°C evaporator temperature. °C Line 5. Thermophysical properties and viscosity data based on calculations 20 1.357 from NIST REFPROP program Version 6.01. 30 1.184 6. For brazed Type L copper tubing larger than 28 mm OD for discharge 40 1.000 or liquid service, see Safety Requirements section. 7. Values are based on 40°C condensing temperature. Multiply table 50 0.801 capacities by the following factors for other condensing temperatures.

Discharge Line 0.765 0.908 1.000 1.021

2014 ASHRAE Handbook—Refrigeration (SI)

Type L Copper, OD, mm 12 15 18 22 28 35 42 54 67 79 105 130 156 206 257 Steel mm SCH 10 80 15 80 20 80 25 80 32 80 40 80 50 40 65 40 80 40 100 40 125 40 150 40 200 40 250 40 300 IDb 350 30 400 30

Suction Lines (t = 0.04 K/m) Saturated Suction Temperature, °C –40 –30 –20 –5 Corresponding p, Pa/m 251.7 350.3 471.6 700.5 0.28 0.44 0.68 1.21 0.53 0.85 1.30 2.31 0.92 1.47 2.26 4.00 1.63 2.60 3.98 7.02 3.22 5.14 7.85 13.83 5.91 9.42 14.37 25.28 9.82 15.65 23.83 41.86 19.55 31.07 47.24 82.83 34.83 55.25 84.08 147.12 54.01 85.61 129.94 227.12 115.54 182.78 277.24 484.29 205.61 325.01 492.45 857.55 333.77 526.96 797.36 1389.26 683.01 1078.30 1631.18 2832.25 1216.78 1916.48 2891.11 5022.65

Halocarbon Refrigeration Systems

1.11

Table 10 Suction Line Capacities in Kilowatts for Refrigerant 22 (Single- or High-Stage Applications) for Pressure Drops of 0.02 and 0.01 K/m Equivalent Nominal Line OD, mm

–40

–30

t = 0.02 p = 97.9

t = 0.01 p = 49.0

t = 0.02 p = 138

12 15 18 22 28 35 42 54 67 79 105

0.21 0.41 0.72 1.28 2.54 4.69 7.82 15.63 27.94 43.43 93.43

0.14 0.28 0.49 0.86 1.72 3.19 5.32 10.66 19.11 29.74 63.99

0.34 0.65 1.13 2.00 3.97 7.32 12.19 24.34 43.48 67.47 144.76

10 15 20 25 32 40 50 65 80 100 125 150 200 250 300

0.33 0.61 1.30 2.46 5.11 7.68 14.85 23.74 42.02 85.84 155.21 251.47 515.37 933.07 1494.35

0.23 0.42 0.90 1.71 3.56 5.36 10.39 16.58 29.43 60.16 108.97 176.49 362.01 656.12 1050.57

0.50 0.94 1.98 3.76 7.79 11.70 22.65 36.15 63.95 130.57 235.58 381.78 781.63 1413.53 2264.54

Saturated Suction Temperature, °C –20 –5 t = 0.01 t = 0.02 t = 0.01 t = 0.02 t = 0.01 p = 69.2 p = 189 p = 94.6 p = 286 p = 143 TYPE L COPPER LINE 0.23 0.51 0.34 0.87 0.59 0.44 0.97 0.66 1.67 1.14 0.76 1.70 1.15 2.91 1.98 1.36 3.00 2.04 5.14 3.50 2.70 5.95 4.06 10.16 6.95 4.99 10.96 7.48 18.69 12.80 8.32 18.20 12.46 31.03 21.27 16.65 36.26 24.88 61.79 42.43 29.76 64.79 44.48 110.05 75.68 46.26 100.51 69.04 170.64 117.39 99.47 215.39 148.34 365.08 251.92 STEEL LINE 0.35 0.74 0.52 1.25 0.87 0.65 1.38 0.96 2.31 1.62 1.38 2.92 2.04 4.87 3.42 2.62 5.52 3.86 9.22 6.47 5.45 11.42 8.01 19.06 13.38 8.19 17.16 12.02 28.60 20.10 14.86 33.17 23.27 55.18 38.83 25.30 52.84 37.13 87.91 61.89 44.84 93.51 65.68 155.62 109.54 91.69 190.95 134.08 317.17 223.47 165.78 344.66 242.47 572.50 403.23 268.72 557.25 391.95 925.72 652.73 550.49 1141.07 803.41 1895.86 1336.79 996.65 2063.66 1454.75 3429.24 2417.91 1593.85 3305.39 2330.50 5477.74 3867.63

5 t = 0.02 p = 366

t = 0.01 p = 183

1.20 2.30 4.00 7.07 13.98 25.66 42.59 84.60 150.80 233.56 499.16

0.82 1.56 2.73 4.82 9.56 17.59 29.21 58.23 103.80 161.10 344.89

1.69 3.15 6.63 12.52 25.88 38.89 74.92 119.37 211.33 430.77 776.67 1255.93 2572.39 4646.48 7433.20

1.18 2.20 4.65 8.79 18.20 27.35 52.77 84.05 148.77 303.17 547.16 885.79 1813.97 3280.83 5248.20

p = pressure drop per unit equivalent line length, Pa/m t = corresponding change in saturation temperature, K/m

Table 11

Nominal Line OD, mm

Suction Line Capacities in Kilowatts for Refrigerant 134a (Single- or High-Stage Applications) for Pressure Drops of 0.02 and 0.01 K/m Equivalent –10

–5

t = 0.02 p = 159

t = 0.01 p = 79.3

t = 0.02 p = 185

12 15 18 22 28 35 42 54 67 79 105

0.42 0.81 1.40 2.48 4.91 9.05 15.00 30.00 53.40 82.80 178.00

0.28 0.55 0.96 1.69 3.36 6.18 10.30 20.50 36.70 56.90 122.00

0.52 0.99 1.73 3.05 6.03 11.10 18.40 36.70 65.40 101.00 217.00

10 15 20 25 32 40 50 65 80 100

0.61 1.13 2.39 4.53 9.37 14.10 27.20 43.30 76.60 156.00

0.42 0.79 1.67 3.17 6.57 9.86 19.10 30.40 53.80 110.00

0.74 1.38 2.91 5.49 11.40 17.10 32.90 52.50 92.80 189.00

p = pressure drop per unit equivalent line length, Pa/m t = corresponding change in saturation temperature, K/m

Saturated Suction Temperature, °C 0 5 t = 0.01 t = 0.02 t = 0.01 t = 0.02 t = 0.01 p = 92.4 p = 212 p = 106 p = 243 p = 121 TYPE L COPPER LINE 0.35 0.63 0.43 0.76 0.51 0.67 1.20 0.82 1.45 0.99 1.18 2.09 1.43 2.53 1.72 2.08 3.69 2.52 4.46 3.04 4.13 7.31 5.01 8.81 6.02 7.60 13.40 9.21 16.20 11.10 12.60 22.30 15.30 26.90 18.40 25.20 44.40 30.50 53.40 36.70 44.90 79.00 54.40 95.00 65.40 69.70 122.00 84.30 147.00 101.00 149.00 262.00 181.00 315.00 217.00 STEEL LINE 0.52 0.89 0.62 1.06 0.74 0.96 1.65 1.16 1.97 1.38 2.03 3.49 2.44 4.17 2.92 3.85 6.59 4.62 7.86 5.52 7.97 13.60 9.57 16.30 11.40 12.00 20.50 14.40 24.40 17.10 23.10 39.50 27.70 47.00 33.10 36.90 62.90 44.30 75.00 52.70 65.30 111.00 78.30 133.00 93.10 133.00 227.00 160.00 270.00 190.00

10 t = 0.02 p = 278

t = 0.01 p = 139

0.91 1.74 3.03 5.34 10.60 19.40 32.10 63.80 113.00 176.00 375.00

0.62 1.19 2.07 3.66 7.24 13.30 22.10 44.00 78.30 122.00 260.00

1.27 2.35 4.94 9.33 19.30 28.90 55.80 88.80 157.00 320.00

0.89 1.65 3.47 6.56 13.60 20.40 39.40 62.70 111.00 226.00

1.12

Table 12 Suction Line Capacities in Kilowatts for Refrigerant 404A (Single- or High-Stage Applications) Line Size Type L Copper, t = 0.02 OD, mm p = 82.7 12 0.11 15 0.21 18 0.36 22 0.64 28 1.27 35 2.34 42 3.90 54 7.79 67 13.93 79 21.63 105 46.52 130 82.96 156 135.08 206 277.62 257 494.78

–50 t = 0.01 p = 41.4

Saturated Suction Temperature, °C –30 –20

–40 t = 0.005 t = 0.02 t = 0.01 p = 20.7 p = 120.3 p = 60.2

t = 0.005 t = 0.02 t = 0.01 p = 30.1 p = 168.6 p = 84.3

–5

5

t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 p = 42.1 p = 227.5 p = 113.8 p = 56.9 p = 339.6 p = 169.8 p = 84.9 p = 431.6 p = 215.8 p = 107.9

0.07 0.14 0.24 0.43 0.86 1.60 2.66 5.33 9.54 14.83 31.94 57.04 92.93 191.32 341.54

0.05 0.09 0.16 0.29 0.59 1.09 1.81 3.63 6.51 10.14 21.86 39.08 63.76 131.43 235.13

0.18 0.35 0.61 1.08 2.15 3.96 6.58 13.14 23.46 36.32 77.93 138.94 225.72 463.83 825.49

0.12 0.24 0.42 0.74 1.47 2.71 4.50 8.99 16.06 24.95 53.67 95.71 155.85 319.80 570.75

0.08 0.16 0.28 0.50 1.00 1.84 3.07 6.15 11.01 17.11 36.81 65.73 107.05 220.68 394.05

0.30 0.57 0.99 1.75 3.46 6.36 10.56 21.01 37.48 58.00 124.23 221.02 359.48 736.05 1309.80

0.20 0.39 0.67 1.19 2.36 4.35 7.24 14.42 25.74 39.98 85.76 152.62 248.41 509.42 908.43

0.14 0.26 0.46 0.81 1.61 2.97 4.94 9.88 17.67 27.42 58.97 105.16 171.22 351.91 626.94

0.46 0.87 1.52 2.68 5.31 9.73 16.16 32.10 57.15 88.53 189.26 336.45 546.48 1116.88 1986.58

0.31 0.60 1.04 1.83 3.63 6.68 11.11 22.10 39.39 61.05 130.73 232.53 378.05 774.88 1380.30

0.21 0.40 0.71 1.25 2.48 4.57 7.60 15.16 27.10 42.02 90.09 160.63 261.35 535.76 955.39

0.82 1.56 2.71 4.77 9.42 17.24 28.59 56.67 100.86 155.88 332.59 590.71 956.48 1956.60 3468.26

0.56 1.07 1.86 3.27 6.47 11.88 19.71 39.12 69.65 107.91 230.75 409.49 665.13 1359.50 2418.47

0.38 0.73 1.27 2.24 4.43 8.15 13.53 26.96 48.01 74.41 159.39 283.55 461.14 942.94 1680.42

1.16 2.21 3.84 6.74 13.28 24.34 40.27 79.78 141.88 218.97 466.69 827.79 1340.68 2738.52 4855.54

0.79 1.51 2.63 4.64 9.15 16.79 27.83 55.22 98.26 151.78 324.29 575.86 934.05 1906.18 3385.31

0.54 1.03 1.80 3.18 6.28 11.53 19.13 38.07 67.87 104.92 224.68 399.34 648.51 1327.72 2355.91

0.07 0.15 0.34 0.66 1.44 2.20 5.14 8.20 14.53 29.72 53.71 87.00 178.72 323.52 518.07 670.58 967.52

0.05 0.10 0.23 0.46 1.00 1.53 3.58 5.73 10.16 20.82 37.65 61.14 125.53 227.79 364.71 472.04 682.10

0.18 0.35 0.80 1.58 3.39 5.18 12.06 19.25 34.02 69.40 125.23 202.96 415.04 751.60 1202.25 1556.41 2242.69

0.12 0.25 0.56 1.10 2.38 3.63 8.47 13.50 23.91 48.87 88.15 142.86 292.83 530.24 847.92 1097.86 1584.19

0.09 0.17 0.39 0.77 1.65 2.53 5.91 9.45 16.74 34.23 61.96 100.36 205.95 372.76 597.02 772.66 1114.80

0.28 0.56 1.26 2.48 5.32 8.12 18.88 30.08 53.25 108.52 195.88 316.73 647.78 1173.25 1874.13 2426.35 3496.21

0.20 0.39 0.88 1.74 3.74 5.69 13.28 21.14 37.41 76.38 137.88 223.39 456.97 827.24 1323.22 1713.06 2468.35

0.14 0.27 0.62 1.21 2.61 3.99 9.30 14.86 26.28 53.78 97.02 157.23 322.29 583.58 933.37 1208.28 1743.57

0.43 0.84 1.90 3.72 7.99 12.18 28.31 45.11 79.70 162.46 293.27 474.25 970.08 1754.74 2807.26 3629.13 5229.67

0.30 0.59 1.33 2.61 5.62 8.56 19.94 31.77 56.12 114.60 206.86 334.48 684.08 1239.01 1979.12 2562.28 3692.16

0.21 0.41 0.93 1.83 3.94 6.01 13.99 22.33 39.52 80.69 145.62 235.68 482.62 874.01 1398.01 1809.93 2607.92

0.74 1.46 3.28 6.43 13.79 21.04 48.83 77.74 137.36 279.72 505.03 816.77 1670.96 3018.58 4822.19 6243.50 8997.43

0.52 1.02 2.31 4.53 9.72 14.82 34.39 54.80 96.81 197.33 356.22 576.04 1178.30 2131.38 3409.82 4408.09 6362.16

0.36 1.03 0.72 2.03 1.62 4.57 3.18 8.95 6.82 19.16 10.41 29.23 24.22 67.87 38.59 107.94 68.30 190.74 139.20 388.91 251.26 700.50 406.28 1132.90 830.92 2317.71 1504.95 4186.92 2403.97 6698.68 3113.72 8673.32 4484.65 12479.89

0.72 1.43 3.21 6.30 13.50 20.59 47.80 76.17 134.57 274.35 494.71 800.04 1636.61 2960.60 4729.58 6123.59 8824.64

0.51 1.00 2.26 4.43 9.52 14.51 33.73 53.70 94.95 193.54 349.38 564.98 1155.67 2090.45 3344.33 4323.43 6230.18

Steel mm SCH 80 0.11 80 0.21 80 0.49 80 0.96 80 2.06 80 3.15 40 7.33 40 11.72 40 20.73 40 42.43 40 76.50 40 123.97 40 254.09 40 460.09 ID* 735.84 30 952.57 30 1374.81

Notes: 1. t = corresponding change in saturation temperature, K/m. 2. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and saturated evaporator outlet temperature. Liquid capacity (kW) based on –5°C evaporator temperature. 3. Thermophysical properties and viscosity data based on calculations from NIST REFPROP program Version 6.01. 4. Values are based on 40°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures. *Pipe inside diameter is same as nominal pipe size.

Condensing Temperature, °C

Suction Line

Discharge Line

20 30 40 50

1.344 1.177 1.000 0.809

0.812 0.906 1.000 1.035

2014 ASHRAE Handbook—Refrigeration (SI)

10 15 20 25 32 40 50 65 80 100 125 150 200 250 300 350 400

Line Size Type L Copper, t = 0.02 OD, mm p = 86.9 12 0.11 15 0.21 18 0.37 22 0.66 28 1.31 35 2.41 42 4.01 54 8.00 67 14.30 79 22.22 105 47.74 130 85.15 156 138.40 206 284.60 257 506.84

–50 t = 0.01 p = 43.4

Saturated Suction Temperature, °C –30 –20

–40 t = 0.005 t = 0.02 t = 0.01 p = 21.7 p = 125.8 p = 62.9

t = 0.005 t = 0.02 t = 0.01 p = 31.5 p = 175.1 p = 87.6

–5

5

t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 p = 43.8 p = 235.8 p = 117.9 p = 58.9 p = 350.3 p = 175.1 p = 87.6 p = 441.3 p = 220.6 p = 110.3

0.07 0.14 0.25 0.45 0.89 1.64 2.74 5.47 9.80 15.23 32.78 58.45 95.37 196.00 349.91

0.05 0.10 0.17 0.30 0.60 1.11 1.86 3.73 6.69 10.40 22.46 40.16 65.53 134.88 240.93

0.19 0.36 0.63 1.11 2.20 4.05 6.73 13.43 23.95 37.15 79.72 141.94 230.65 473.07 843.99

0.13 0.24 0.43 0.75 1.50 2.77 4.61 9.20 16.45 25.52 54.80 97.75 159.24 326.88 583.29

0.09 0.16 0.29 0.51 1.02 1.88 3.15 6.29 11.26 17.49 37.64 67.24 109.54 225.50 402.74

0.30 0.58 1.01 1.78 3.52 6.47 10.76 21.41 38.13 59.12 126.44 225.05 365.21 748.74 1331.07

0.20 0.39 0.69 1.21 2.41 4.43 7.37 14.70 26.22 40.66 87.27 155.37 252.88 518.73 923.35

0.14 0.27 0.47 0.83 1.64 3.03 5.04 10.06 18.00 27.93 60.03 107.07 174.27 358.26 638.46

0.46 0.89 1.55 2.73 5.39 9.89 16.40 32.61 58.06 89.96 192.36 341.35 554.50 1133.28 2016.20

0.32 0.61 1.06 1.86 3.69 6.79 11.28 22.44 40.01 62.02 132.88 236.40 384.41 786.17 1400.80

0.21 0.41 0.72 1.27 2.53 4.65 7.73 15.42 27.53 42.68 91.53 163.23 265.63 544.81 969.79

0.83 1.58 2.75 4.83 9.54 17.46 28.95 57.40 102.02 157.67 336.46 597.65 967.77 1975.61 3509.62

0.57 1.08 1.88 3.31 6.55 12.03 19.95 39.62 70.56 109.13 233.43 414.27 673.08 1375.65 2447.80

0.38 0.74 1.29 2.27 4.49 8.26 13.71 27.30 48.63 75.38 161.52 286.80 466.52 953.94 1700.43

1.17 2.23 3.87 6.80 13.39 24.50 40.60 80.32 142.60 220.49 469.99 833.69 1350.40 2752.34 4885.91

0.80 1.53 2.65 4.67 9.22 16.89 28.00 55.58 98.93 152.82 326.56 579.94 940.74 1920.12 3410.38

0.54 1.04 1.82 3.20 6.33 11.62 19.28 38.32 68.33 105.83 226.24 402.13 653.12 1334.18 2373.45

0.08 0.15 0.34 0.68 1.47 2.25 5.26 8.40 14.87 30.41 54.98 89.07 182.50 330.97 529.04 685.81 990.11

0.05 0.10 0.24 0.47 1.02 1.56 3.66 5.86 10.39 21.28 38.55 62.50 128.37 232.63 372.46 482.78 696.48

0.18 0.36 0.82 1.61 3.46 5.28 12.30 19.59 34.68 70.82 127.81 206.64 423.13 765.23 1225.83 1584.62 2286.84

0.13 0.25 0.57 1.12 2.42 3.70 8.62 13.77 24.37 49.75 89.98 145.43 298.14 539.85 864.69 1117.74 1612.91

0.09 0.17 0.40 0.78 1.69 2.58 6.03 9.64 17.09 34.94 63.16 102.33 209.65 380.04 607.74 787.98 1136.96

0.29 0.57 1.28 2.51 5.40 8.24 19.15 30.51 54.00 110.06 198.66 321.22 657.85 1189.87 1903.47 2464.42 3545.81

0.20 0.40 0.90 1.76 3.79 5.79 13.46 21.48 38.02 77.58 140.00 226.55 463.90 838.96 1342.41 1737.32 2507.22

0.14 0.28 0.63 1.23 2.65 4.06 9.45 15.07 26.72 54.55 98.51 159.43 326.85 591.84 946.41 1225.41 1768.24

0.43 0.85 1.92 3.77 8.10 12.35 28.67 45.68 80.71 164.51 296.97 480.23 982.32 1776.88 2842.68 3674.91 5304.13

0.30 0.60 1.35 2.65 5.69 8.67 20.19 32.17 56.94 116.05 209.47 338.70 692.92 1254.64 2004.70 2595.39 3738.73

0.21 0.42 0.94 1.85 3.99 6.09 14.20 22.61 40.01 81.71 147.43 238.99 488.63 885.04 1415.67 1832.75 2640.86

0.75 1.47 3.31 6.49 13.92 21.24 49.30 78.58 138.85 282.44 509.92 824.68 1687.15 3047.81 4868.99 6303.96 9084.58

0.52 1.03 2.33 4.57 9.81 14.96 34.72 55.33 97.85 199.48 359.67 581.62 1189.71 2152.06 3442.84 4450.78 6423.78

0.37 1.03 0.72 2.04 1.64 4.59 3.21 9.00 6.90 19.29 10.52 29.38 24.46 68.21 38.97 108.49 68.96 191.71 140.55 390.88 253.70 704.04 410.21 1138.62 840.10 2329.43 1519.53 4208.08 2430.82 6732.54 3147.18 8717.17 4528.18 12542.98

0.73 1.43 3.23 6.33 13.59 20.70 48.14 76.55 135.25 275.74 497.22 804.08 1644.89 2975.57 4753.49 6154.55 8869.25

0.51 1.01 2.27 4.46 9.56 14.59 33.90 54.02 95.43 194.52 351.14 567.84 1161.52 2101.01 3361.24 4345.29 6261.68

Halocarbon Refrigeration Systems

Table 13 Suction Line Capacities in Kilowatts for Refrigerant 507A (Single- or High-Stage Applications)

Steel mm SCH 10 15 20 25 32 40 50 65 80 100 125 150 200 250 300 350 400

80 0.11 80 0.22 80 0.50 80 0.98 80 2.11 80 3.22 40 7.50 40 11.97 40 21.21 40 43.30 40 78.33 40 126.59 40 259.48 40 469.84 ID* 751.43 30 972.76 30 1403.68

Condensing Temperature, °C

Suction Line

Discharge Line

20 30 40 50

1.357 1.184 1.000 0.801

0.765 0.908 1.000 1.021

1.13

Notes: 1. t = corresponding change in saturation temperature, K/m. 2. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and saturated evaporator outlet temperature. Liquid capacity (kW) based on –5°C evaporator temperature. 3. Thermophysical properties and viscosity data based on calculations from NIST REFPROP program Version 6.01. 4. Values are based on 40°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures. *Pipe inside diameter is same as nominal pipe size.

1.14

Table 14 Suction Line Capacities in Kilowatts for Refrigerant 410A (Single- or High-Stage Applications) Line Size Saturated Suction Temperature, °C –50 –40 –30 –20 Type L Copper, t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 t = 0.02 OD, mm p = 109.3 p = 54.6 p = 27.3 p = 158.6 p = 79.3 p = 39.6 p = 221.7 p = 110.8 p = 55.4 p = 299.6 p = 149.8 p = 74.9 p = 447.1 12 0.21 0.14 0.10 0.35 0.24 0.16 0.55 0.37 0.25 0.82 0.56 0.38 1.40 15 0.41 0.28 0.19 0.67 0.46 0.31 1.05 0.71 0.48 1.57 1.07 0.72 2.68 18 0.72 0.49 0.33 1.17 0.80 0.54 1.83 1.24 0.84 2.73 1.86 1.27 4.65 22 1.27 0.86 0.59 2.08 1.41 0.96 3.22 2.20 1.50 4.81 3.28 2.24 8.19 28 2.54 1.73 1.17 4.12 2.81 1.91 6.39 4.37 2.98 9.51 6.51 4.44 16.15 35 4.67 3.19 2.16 7.59 5.18 3.53 11.75 8.04 5.50 17.44 11.95 8.18 29.56 42 7.79 5.31 3.62 12.63 8.63 5.89 19.50 13.37 9.15 28.92 19.88 13.61 49.03 54 15.56 10.63 7.25 25.15 17.23 11.77 38.82 26.67 18.26 57.48 39.55 27.17 97.22 67 27.80 19.04 12.99 44.92 30.80 21.09 69.13 47.55 32.64 102.34 70.53 48.52 172.78 79 43.12 29.57 20.23 69.55 47.85 32.76 107.18 73.87 50.66 158.27 109.33 75.23 267.04 105 92.76 63.67 43.59 149.23 102.78 70.55 229.65 158.20 108.97 338.41 234.20 161.33 569.83 130 165.44 113.74 77.94 265.67 183.33 125.87 408.65 282.18 194.35 601.64 416.66 287.70 1011.16 156 268.90 185.37 127.16 432.51 297.99 205.31 663.37 458.32 316.51 977.30 676.05 467.21 1638.88 206 552.84 381.00 262.20 887.06 612.75 422.22 1358.62 940.09 649.28 1997.41 1385.82 959.17 3345.90 257 986.70 680.26 468.27 1578.79 1091.59 753.27 2417.55 1676.97 1159.08 3554.05 2468.86 1708.47 5943.90

–5

5

t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 p = 223.6 p = 111.8 p = 568.8 p = 284.4 p = 142.2

0.96 1.83 3.19 5.61 11.09 20.38 33.75 67.10 119.50 184.82 395.31 701.60 1139.70 2329.79 4140.36

0.65 1.25 2.17 3.84 7.61 13.99 23.19 46.17 82.35 127.67 273.54 485.69 790.11 1615.84 2879.83

1.94 3.69 6.41 11.26 22.19 40.66 67.28 133.10 236.73 365.38 778.82 1381.55 2237.78 4560.98 8106.26

1.32 2.53 4.39 7.74 15.28 27.99 46.41 92.11 163.91 253.23 541.15 961.03 1558.94 3182.24 5651.46

0.90 1.73 3.01 5.31 10.50 19.28 31.96 63.61 113.23 175.38 374.91 666.39 1082.28 2210.93 3933.16

Steel mm SCH 80 0.21 80 0.43 80 0.97 80 1.90 80 4.10 80 6.27 40 14.60 40 23.29 40 41.31 40 84.29 40 152.27 40 246.40 40 505.05 40 914.47 ID* 1462.58 30 1893.35 30 2732.08

0.15 0.29 0.67 1.32 2.86 4.37 10.22 16.31 28.96 59.20 107.03 173.39 355.23 643.03 1029.72 1335.27 1923.07

0.10 0.20 0.46 0.92 1.99 3.04 7.13 11.41 20.23 41.43 75.04 121.50 249.51 452.76 724.90 938.33 1353.65

0.34 0.68 1.53 3.01 6.48 9.89 23.03 36.74 65.02 132.64 239.36 387.47 793.40 1434.85 2298.53 2971.29 4288.04

0.24 0.47 1.07 2.11 4.53 6.94 16.16 25.77 45.70 93.25 168.49 272.69 558.74 1011.71 1618.72 2094.69 3022.68

0.16 0.33 0.74 1.47 3.16 4.84 11.32 18.08 32.04 65.51 118.43 191.87 393.11 712.74 1139.56 1477.51 2131.16

0.52 1.03 2.33 4.57 9.81 14.96 34.78 55.41 98.08 199.89 360.80 583.40 1194.79 2161.06 3457.11 4475.91 6439.96

0.36 0.72 1.63 3.20 6.88 10.51 24.45 39.02 69.05 140.72 254.27 411.47 842.54 1523.73 2438.12 3155.35 4548.16

0.25 0.50 1.13 2.23 4.82 7.36 17.16 27.37 48.53 99.07 178.92 289.61 593.64 1074.91 1718.90 2225.57 3211.53

0.76 1.50 3.39 6.65 14.28 21.74 50.53 80.52 142.24 289.94 523.41 846.41 1731.34 3131.74 5010.22 6477.03 9333.77

0.53 1.05 2.38 4.67 10.02 15.29 35.59 56.70 100.36 204.54 369.19 596.96 1220.92 2211.30 3532.27 4573.09 6589.52

0.37 1.27 0.89 0.73 2.49 1.75 1.66 5.61 3.95 3.27 11.00 7.74 7.03 23.58 16.61 10.73 35.97 25.34 25.03 83.50 58.81 39.85 133.08 93.70 70.52 235.15 165.75 144.02 478.33 337.88 259.89 863.60 609.14 421.22 1396.67 985.04 861.37 2857.36 2014.90 1559.90 5161.78 3644.74 2495.13 8246.13 5830.80 3230.24 10 676.40 7537.85 4654.45 15 385.64 10 879.32

Notes: 1. t = corresponding change in saturation temperature, K/m. 2. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and saturated evaporator outlet temperature. Liquid capacity (kW) based on –5°C evaporator temperature. 3. Thermophysical properties and viscosity data based on calculations from NIST REFPROP program Version 6.01. 4. Values are based on 40°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures. *Pipe inside diameter is same as nominal pipe size.

0.62 1.71 1.21 0.84 1.23 3.38 2.38 1.67 2.77 7.61 5.35 3.76 5.44 14.91 10.49 7.38 11.67 31.97 22.53 15.84 17.82 48.69 34.30 24.16 41.43 113.04 79.77 56.18 66.01 179.78 126.86 89.51 116.80 317.68 224.14 158.14 238.04 647.74 456.93 322.34 429.66 1166.69 823.96 581.89 694.73 1886.85 1332.47 940.98 1422.80 3860.17 2725.80 1924.78 2573.47 6973.35 4930.92 3481.66 4116.85 11 156.71 7877.16 5570.02 5330.07 14 445.51 10 198.90 7200.72 7668.93 20 785.39 14 697.52 10 376.43

Condensing Temperature, °C

Suction Line

Discharge Line

20 30 40 50

1.238 1.122 1.000 0.867

0.657 0.866 1.000 1.117

2014 ASHRAE Handbook—Refrigeration (SI)

10 15 20 25 32 40 50 65 80 100 125 150 200 250 300 350 400

Line Size Type L Copper, t = 0.02 OD, mm p = 56.8 12 0.09 15 0.18 18 0.32 22 0.57 28 1.14 35 2.10 42 3.51 54 7.04 67 12.60 79 19.60 105 42.25 130 75.52 156 123.08 206 253.51 257 453.04

–50

Saturated Suction Temperature, °C –30 –20

–40

t = 0.01 p = 28.4

t = 0.005 p = 14.2

t = 0.02 p = 86.9

t = 0.01 p = 43.4

t = 0.005 t = 0.02 t = 0.01 p = 21.7 p = 127.2 p = 63.6

t = 0.005 t = 0.02 t = 0.01 p = 31.8 p = 179.3 p = 89.6

0.06 0.12 0.22 0.38 0.77 1.43 2.39 4.79 8.60 13.40 28.89 51.75 84.50 174.21 311.95

0.04 0.08 0.15 0.26 0.52 0.97 1.62 3.26 5.85 9.13 19.73 35.35 57.78 119.60 214.12

0.17 0.32 0.56 1.00 1.99 3.67 6.11 12.22 21.84 33.95 72.93 130.18 212.26 435.89 777.76

0.11 0.22 0.38 0.68 1.35 2.50 4.17 8.34 14.93 23.24 50.07 89.49 145.88 300.07 536.42

0.08 0.15 0.26 0.46 0.91 1.69 2.83 5.69 10.22 15.89 34.27 61.30 100.08 206.07 368.41

0.28 0.53 0.93 1.65 3.28 6.05 10.06 20.07 35.86 55.68 119.35 212.71 346.03 710.91 1267.28

0.19 0.36 0.63 1.12 2.24 4.13 6.87 13.74 24.58 38.15 82.12 146.43 238.65 490.35 874.79

0.13 0.24 0.43 0.76 1.52 2.81 4.68 9.38 16.80 26.15 56.27 100.59 164.01 337.87 603.28

0.44 0.85 1.48 2.61 5.17 9.52 15.82 31.50 56.25 87.17 186.59 331.83 539.39 1106.41 1967.94

0.07 0.14 0.31 0.61 1.33 2.03 4.78 7.64 13.57 27.82 50.49 81.94 168.19 305.12 489.17 632.96 914.52

0.05 0.09 0.21 0.42 0.92 1.41 3.31 5.29 9.44 19.34 35.20 57.09 117.67 213.63 342.30 443.54 641.60

0.17 0.34 0.76 1.51 3.25 4.96 11.60 18.51 32.80 67.09 121.04 196.05 401.96 727.74 1165.00 1508.75 2173.71

0.12 0.23 0.53 1.05 2.26 3.46 8.10 12.95 22.95 47.03 84.98 137.62 282.75 512.49 819.31 1062.23 1532.53

0.08 0.16 0.37 0.72 1.57 2.41 5.64 9.01 16.01 32.88 59.53 96.49 198.35 359.86 575.99 746.66 1077.06

0.28 0.54 1.23 2.43 5.22 7.98 18.59 29.65 52.51 107.23 193.47 312.76 641.19 1161.16 1854.62 2404.53 3464.70

0.19 0.38 0.86 1.69 3.65 5.58 13.03 20.77 36.82 75.32 136.02 220.12 451.12 818.01 1308.26 1693.25 2443.72

0.13 0.26 0.60 1.18 2.54 3.89 9.11 14.53 25.82 52.77 95.36 154.67 317.28 575.07 920.85 1192.13 1719.98

0.43 0.84 1.90 3.74 8.03 12.27 28.50 45.41 80.36 163.95 295.91 479.06 979.77 1774.57 2834.60 3669.84 5288.11

–5

5

t = 0.005 p = 44.8

t = 0.02 p = 281

t = 0.01 t = 0.005 t = 0.02 t = 0.01 t = 0.005 p = 140.5 p = 70.2 p = 367.1 p = 183.6 p = 91.8

0.30 0.58 1.00 1.78 3.53 6.51 10.84 21.63 38.59 59.91 128.63 229.25 372.95 765.96 1362.93

0.20 0.39 0.68 1.21 2.41 4.44 7.41 14.81 26.45 41.09 88.40 157.72 256.85 528.25 942.32

0.81 1.56 2.71 4.78 9.45 17.36 28.75 57.14 101.74 157.62 337.06 598.18 971.60 1985.99 3532.80

0.55 1.06 1.85 3.26 6.47 11.90 19.76 39.32 70.12 108.68 232.78 414.13 673.43 1377.71 2454.46

0.38 0.72 1.26 2.23 4.43 8.14 13.55 27.02 48.24 74.81 160.43 286.04 465.49 954.72 1698.01

0.80 1.53 2.66 4.70 9.29 17.06 28.32 56.27 100.33 155.40 332.20 590.48 957.92 1959.07 3484.73

0.54 1.04 1.82 3.21 6.36 11.70 19.44 38.73 69.03 106.97 229.48 408.14 663.50 1359.63 2421.38

0.30 0.59 1.33 2.61 5.63 8.60 20.04 31.91 56.59 115.43 208.47 337.13 690.92 1251.23 1999.14 2591.52 3733.44

0.21 0.41 0.93 1.82 3.93 6.02 14.04 22.38 39.68 81.16 146.56 237.18 486.09 881.14 1409.65 1824.48 2633.11

0.76 1.49 3.37 6.61 14.19 21.65 50.24 80.05 141.41 288.25 520.35 841.46 1721.21 3113.43 4980.90 6439.13 9295.35

0.53 1.05 2.36 4.64 9.97 15.20 35.38 56.36 99.77 203.34 367.03 593.47 1213.78 2198.37 3511.61 4546.34 6550.98

0.37 1.07 0.75 0.73 2.11 1.48 1.65 4.74 3.33 3.25 9.29 6.54 6.99 19.96 14.04 10.67 30.39 21.41 24.87 70.68 49.78 39.62 112.40 79.31 70.11 198.60 140.12 143.12 404.88 285.62 258.39 730.09 515.60 418.57 1180.67 833.78 856.37 2415.26 1705.51 1550.04 4369.17 3085.01 2479.36 6979.76 4935.46 3209.82 9037.00 6380.38 4627.41 13 023.12 9194.30

0.52 1.03 2.34 4.59 9.86 15.06 35.00 55.80 98.69 201.57 363.33 588.29 1202.68 2178.30 3479.49 4504.75 6491.20

1.17 2.24 3.89 6.86 13.54 24.80 41.12 81.62 145.23 224.39 479.50 850.78 1378.45 2819.41 5007.64

Halocarbon Refrigeration Systems

Table 15 Suction Line Capacities in Kilowatts for Refrigerant 407C (Single- or High-Stage Applications)

Steel mm SCH 10 80 0.10 15 80 0.20 20 80 0.45 25 80 0.89 32 80 1.92 40 80 2.93 50 40 6.87 65 40 10.97 80 40 19.46 100 40 39.85 125 40 72.01 150 40 116.91 200 40 239.49 250 40 434.62 300 ID* 695.90 350 30 900.71 400 30 1299.44

Condensing Temperature, °C

Suction Line

Discharge Line

20 30 40 50

1.202 1.103 1.000 0.891

0.605 0.845 1.000 1.133

1.15

Notes: 1. t = corresponding change in saturation temperature, K/m. 2. Capacity (kW) based on standard refrigerant cycle of 40°C liquid and saturated evaporator outlet temperature. Liquid capacity (kW) based on –5°C evaporator temperature. 3. Thermophysical properties and viscosity data based on calculations from NIST REFPROP program Version 6.01. 4. Values are based on 40°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures. *Pipe inside diameter is same as nominal pipe size.

1.16

2014 ASHRAE Handbook—Refrigeration (SI) One of the most difficult problems in low-temperature refrigeration systems using halocarbon refrigerants is returning lubrication oil from the evaporator to the compressors. Except for most centrifugal compressors and rarely used nonlubricated compressors, refrigerant continuously carries oil into the discharge line from the compressor. Most of this oil can be removed from the stream by an oil separator and returned to the compressor. Coalescing oil separators are far better than separators using only mist pads or baffles; however, they are not 100% effective. Oil that finds its way into the system must be managed. Oil mixes well with halocarbon refrigerants at higher temperatures. As temperature decreases, miscibility is reduced, and some oil separates to form an oil-rich layer near the top of the liquid level in a flooded evaporator. If the temperature is very low, the oil becomes a gummy mass that prevents refrigerant controls from functioning, blocks flow passages, and fouls heat transfer surfaces. Proper oil management is often key to a properly functioning system. In general, direct-expansion and liquid overfeed system evaporators have fewer oil return problems than do flooded system evaporators because refrigerant flows continuously at velocities high enough to sweep oil from the evaporator. Low-temperature systems using hot-gas defrost can also be designed to sweep oil out of the circuit each time the system defrosts. This reduces the possibility of oil coating the evaporator surface and hindering heat transfer. Flooded evaporators can promote oil contamination of the evaporator charge because they may only return dry refrigerant vapor back to the system. Skimming systems must sample the oil-rich layer floating in the drum, a heat source must distill the refrigerant, and the oil must be returned to the compressor. Because flooded halocarbon systems can be elaborate, some designers avoid them. System Capacity Reduction. Using automatic capacity control on compressors requires careful analysis and design. The compressor can load and unload as it modulates with system load requirements

Example 2. Determine the line size and pressure drop equivalent (in degrees) for the suction line of a 105 kW R-22 system, operating at 5°C suction and 40°C condensing temperatures. Suction line is copper tubing, with 15 m of straight pipe and six long-radius elbows. Solution: Add 50% to the straight length of pipe to establish a trial equivalent length. Trial equivalent length is 15  1.5 = 22.5 m. From Table 3 (for 5°C suction, 40°C condensing), 122.7 kW capacity in 54 mm OD results in a 0.04 K loss per metre equivalent length. Straight pipe length Six 50 mm long-radius elbows at 1.0 m each (Table 16)

= =

15.0 m 6.0 m

Total equivalent length

=

21.0 m

t = 0.0421.0(105/122.7)1.8 = 0.63 K Because 0.63 K is below the recommended 1 K, recompute for the next smaller (42 mm) tube (i.e., t = 2.05 K). This temperature drop is too large; therefore, the 54 mm tube is recommended.

Oil Management in Refrigerant Lines Oil Circulation. All compressors lose some lubricating oil during normal operation. Because oil inevitably leaves the compressor with the discharge gas, systems using halocarbon refrigerants must return this oil at the same rate at which it leaves (Cooper 1971). Oil that leaves the compressor or oil separator reaches the condenser and dissolves in the liquid refrigerant, enabling it to pass readily through the liquid line to the evaporator. In the evaporator, the refrigerant evaporates, and the liquid phase becomes enriched in oil. The concentration of refrigerant in the oil depends on the evaporator temperature and types of refrigerant and oil used. The viscosity of the oil/refrigerant solution is determined by the system parameters. Oil separated in the evaporator is returned to the compressor by gravity or by drag forces of the returning gas. Oil’s effect on pressure drop is large, increasing the pressure drop by as much as a factor of 10 (Alofs et al. 1990).

Table 16 Fitting Losses in Equivalent Metres of Pipe (Screwed, Welded, Flanged, Flared, and Brazed Connections) Smooth Bend Elbows 90° Stda

90° LongRadiusb

90° Streeta

45° Stda

Smooth Bend Tees 45° Streeta

180° Stda

Flow Through Branch

Nominal Pipe or Tube Size, mm 10 15 20 25 32 40 50 65 80 90 100 125 150 200 250 300 350 400 450 500 600 a R/D

0.4 0.5 0.6 0.8 1.0 1.2 1.5 1.8 2.3 2.7 3.0 4.0 4.9 6.1 7.6 9.1 10 12 13 15 18

approximately equal to 1.

0.3 0.3 0.4 0.5 0.7 0.8 1.0 1.2 1.5 1.8 2.0 2.5 3.0 4.0 4.9 5.8 7.0 7.9 8.8 10 12

0.7 0.8 1.0 1.2 1.7 1.9 2.5 3.0 3.7 4.6 5.2 6.4 7.6 — — — — — — — — b R/D

0.2 0.2 0.3 0.4 0.5 0.6 0.8 1.0 1.2 1.4 1.6 2.0 2.4 3.0 4.0 4.9 5.5 6.1 7.0 7.9 9.1

approximately equal to 1.5.

0.3 0.4 0.5 0.6 0.9 1.0 1.4 1.6 2.0 2.2 2.6 3.4 4.0 — — — — — — — —

0.7 0.8 1.0 1.2 1.7 1.9 2.5 3.0 3.7 4.6 5.2 6.4 7.6 10 13 15 17 19 21 25 29

0.8 0.9 1.2 1.5 2.1 2.4 3.0 3.7 4.6 5.5 6.4 7.6 9 12 15 18 21 24 26 30 35

Straight-Through Flow No Reduction

Reduced 1/4

Reduced 1/2

0.3 0.3 0.4 0.5 0.7 0.8 1.0 1.2 1.5 1.8 2.0 2.5 3.0 4.0 4.9 5.8 7.0 7.9 8.8 10 12

0.4 0.4 0.6 0.7 0.9 1.1 1.4 1.7 2.1 2.4 2.7 3.7 4.3 5.5 7.0 7.9 9.1 11 12 13 15

0.4 0.5 0.6 0.8 1.0 1.2 1.5 1.8 2.3 2.7 3.0 4.0 4.9 6.1 7.6 9.1 10 12 13 15 18

Halocarbon Refrigeration Systems

1.17 Riser Sizing. The following example demonstrates the use of Table 19 in establishing maximum riser sizes for satisfactory oil transport down to minimum partial loading.

through a considerable range of capacity. A single compressor can unload down to 25% of full-load capacity, and multiple compressors connected in parallel can unload to a system capacity of 12.5% or lower. System piping must be designed to return oil at the lowest loading, yet not impose excessive pressure drops in the piping and equipment at full load. Oil Return up Suction Risers. Many refrigeration piping systems contain a suction riser because the evaporator is at a lower level than the compressor. Oil circulating in the system can return up gas risers only by being transported by returning gas or by auxiliary means such as a trap and pump. The minimum conditions for oil transport correlate with buoyancy forces (i.e., density difference between liquid and vapor, and momentum flux of vapor) (Jacobs et al. 1976). The principal criteria determining the transport of oil are gas velocity, gas density, and pipe inside diameter. Density of the oil/ refrigerant mixture plays a somewhat lesser role because it is almost constant over a wide range. In addition, at temperatures somewhat lower than –40°C, oil viscosity may be significant. Greater gas velocities are required as temperature drops and the gas becomes less dense. Higher velocities are also necessary if the pipe diameter increases. Table 19 translates these criteria to minimum refrigeration capacity requirements for oil transport. Suction risers must be sized for minimum system capacity. Oil must be returned to the compressor at the operating condition corresponding to the minimum displacement and minimum suction temperature at which the compressor will operate. When suction or evaporator pressure regulators are used, suction risers must be sized for actual gas conditions in the riser. For a single compressor with capacity control, the minimum capacity is the lowest capacity at which the unit can operate. For multiple compressors with capacity control, the minimum capacity is the lowest at which the last operating compressor can run. Table 17

Example 3. Determine the maximum size suction riser that will transport oil at minimum loading, using R-22 with a 120 kW compressor with capacity in steps of 25, 50, 75, and 100%. Assume the minimum system loading is 30 kW at 5°C suction and 40°C condensing temperatures with 10 K superheat. Solution: From Table 19, a 54 mm OD pipe at 5°C suction and 30°C liquid temperature has a minimum capacity of 23.1 kW. From the chart at the bottom of Table 19, the correction multiplier for 40°C suction temperature is about 1. Therefore, 54 mm OD pipe is suitable.

Based on Table 19, the next smaller line size should be used for marginal suction risers. When vertical riser sizes are reduced to provide satisfactory minimum gas velocities, pressure drop at full load increases considerably; horizontal lines should be sized to keep total pressure drop within practical limits. As long as horizontal lines are level or pitched in the direction of the compressor, oil can be transported with normal design velocities. Because most compressors have multiple capacity-reduction features, gas velocities required to return oil up through vertical suction risers under all load conditions are difficult to maintain. When the suction riser is sized to allow oil return at the minimum operating capacity of the system, pressure drop in this portion of the line may be too great when operating at full load. If a correctly sized suction riser imposes too great a pressure drop at full load, a double suction riser should be used. Oil Return up Suction Risers: Multistage Systems. Oil movement in the suction lines of multistage systems requires the same design approach as that for single-stage systems. For oil to flow up along a pipe wall, a certain minimum drag of gas flow is required.

Special Fitting Losses in Equivalent Metres of Pipe

Sudden Enlargement, d/D

Sudden Contraction, d/D

Sharp Edge

Pipe Projection

1/4

1/2

3/4

1/4

1/2

3/4

Entrance

Exit

Entrance

Exit

10 15 20

0.4 0.5 0.8

0.2 0.3 0.5

0.1 0.1 0.2

0.2 0.3 0.4

0.2 0.3 0.3

0.1 0.1 0.2

0.5 0.5 0.9

0.2 0.3 0.4

0.5 0.5 0.9

0.3 0.5 0.7

25 32 40

1.0 1.4 1.8

0.6 0.9 1.1

0.2 0.3 0.4

0.5 0.7 0.9

0.4 0.5 0.7

0.2 0.3 0.4

1.1 1.6 2.0

0.5 0.8 1.0

1.1 1.6 2.0

0.8 1.3 1.5

50 65 80

2.4 3.0 4.0

1.5 1.9 2.4

0.5 0.6 0.8

1.2 1.5 2.0

0.9 1.2 1.5

0.5 0.6 0.8

2.7 3.7 4.3

1.3 1.7 2.2

2.7 3.7 4.3

2.1 2.7 3.8

90 100 125

4.6 5.2 7.3

2.8 3.4 4.6

0.9 1.2 1.5

2.3 2.7 3.7

1.8 2.1 2.7

0.9 1.2 1.5

5.2 6.1 8.2

2.6 3.0 4.3

5.2 6.1 8.2

4.0 4.9 6.1

150 200 250

8.8 — —

6.7 7.6 9.8

1.8 2.6 3.4

4.6 — —

3.4 4.6 6.1

1.8 2.6 3.4

10 14 18

300 350 400

— — —

12.4 — —

4.0 4.9 5.5

— — —

7.6 — —

4.0 4.9 5.5

22 26 29

450 500 600

— — —

— — —

6.1 — —

— — —

— — —

6.1 — —

35 43 50

Nominal Pipe or Tube Size, mm

Note: Enter table for losses at smallest diameter d.

5.8 7.3 8.8

10 14 18

7.6 10 14

11 14 15

22 26 29

17 20 23

18 21 25

35 43 50

27 33 40

1.18

2014 ASHRAE Handbook—Refrigeration (SI) Table 18 Valve Losses in Equivalent Metres of Pipe

Nominal Pipe or 60° Tube Size, mm Globea Wye 10 15 20 25 32 40 50 65 80 90 100 125 150 200 250 300 350 400 450 500 600

5.2 5.5 6.7 8.8 12 13 17 21 26 30 37 43 52 62 85 98 110 125 140 160 186

2.4 2.7 3.4 4.6 6.1 7.3 9.1 11 13 15 18 22 27 35 44 50 56 64 73 84 98

45° Wye 1.8 2.1 2.1 3.7 4.6 5.5 7.3 8.8 11 13 14 18 21 26 32 40 47 55 61 72 81

Swing Anglea Gateb Checkc 1.8 2.1 2.1 3.7 4.6 5.5 7.3 8.8 11 13 14 18 21 26 32 40 47 55 61 72 81

0.2 0.2 0.3 0.3 0.5 0.5 0.73 0.9 1.0 1.2 1.4 1.8 2.1 2.7 3.7 4.0 4.6 5.2 5.8 6.7 7.6

1.5 1.8 2.2 3.0 4.3 4.9 6.1 7.6 9.1 10 12 15 18 24 30 37 41 46 50 61 73

Lift Check Globe and vertical lift same as globe valved

Fig. 3 Double-Suction Riser Construction

Angle lift same as angle valve

Note: Losses are for valves in fully open position and with screwed, welded, flanged, or flared connections. a These losses do not apply to valves with needlepoint seats. b Regular and short pattern plug cock valves, when fully open, have same loss as gate valve. For valve losses of short pattern plug cocks above 150 mm, check with manufacturer. c Losses also apply to the in-line, ball-type check valve. d For Y pattern globe lift check valve with seat approximately equal to the nominal pipe diameter, use values of 60° wye valve for loss.

Drag can be represented by the friction gradient. The following sizing data may be used for ensuring oil return up vertical suction lines for refrigerants other than those listed in Tables 19 and 20. The line size selected should provide a pressure drop equal to or greater than that shown in the chart. Line Size

Saturation Temperature, °C

50 mm or less

Above 50 mm

–18 –46

80 Pa/m 100 Pa/m

45 Pa/m 57 Pa/m

Double Suction Risers. Figure 3 shows two methods of double suction riser construction. Oil return in this arrangement is accomplished at minimum loads, but it does not cause excessive pressure drops at full load. Sizing and operation of a double suction riser are as follows: 1. Riser A is sized to return oil at minimum load possible. 2. Riser B is sized for satisfactory pressure drop through both risers at full load. The usual method is to size riser B so that the combined cross-sectional area of A and B is equal to or slightly greater than the cross-sectional area of a single pipe sized for acceptable pressure drop at full load without regard for oil return at minimum load. The combined cross-sectional area, however, should not be greater than the cross-sectional area of a single pipe that would return oil in an upflow riser under maximum load. 3. A trap is introduced between the two risers, as shown in both methods. During part-load operation, gas velocity is not sufficient to return oil through both risers, and the trap gradually fills up with oil until riser B is sealed off. The gas then travels up riser A only with enough velocity to carry oil along with it back into the horizontal suction main.

The trap’s oil-holding capacity is limited by close-coupling the fittings at the bottom of the risers. If this is not done, the trap can accumulate enough oil during part-load operation to lower the compressor crankcase oil level. Note in Figure 3 that riser lines A and B form an inverted loop and enter the horizontal suction line from the top. This prevents oil drainage into the risers, which may be idle during part-load operation. The same purpose can be served by running risers horizontally into the main, provided that the main is larger in diameter than either riser Often, double suction risers are essential on low-temperature systems that can tolerate very little pressure drop. Any system using these risers should include a suction trap (accumulator) and a means of returning oil gradually. For systems operating at higher suction temperatures, such as for comfort air conditioning, single suction risers can be sized for oil return at minimum load. Where single compressors are used with capacity control, minimum capacity is usually 25 or 33% of maximum displacement. With this low ratio, pressure drop in single suction risers designed for oil return at minimum load is rarely serious at full load. When multiple compressors are used, one or more may shut down while another continues to operate, and the maximum-tominimum ratio becomes much larger. This may make a double suction riser necessary. The remaining suction line portions are sized to allow a practical pressure drop between the evaporators and compressors because oil is carried along in horizontal lines at relatively low gas velocities. It is good practice to give some pitch to these lines toward the compressor. Avoid traps, but when that is impossible, the risers from them are treated the same as those leading from the evaporators. Preventing Oil Trapping in Idle Evaporators. Suction lines should be designed so that oil from an active evaporator does not drain into an idle one. Figure 4A shows multiple evaporators on different floor levels with the compressor above. Each suction line is brought upward and looped into the top of the common suction line to prevent oil from draining into inactive coils. Figure 4B shows multiple evaporators stacked on the same level, with the compressor above. Oil cannot drain into the lowest evaporator because the common suction line drops below the outlet of the lowest evaporator before entering the suction riser. Figure 4C shows multiple evaporators on the same level, with the compressor located below. The suction line from each evaporator drops down into the common suction line so that oil cannot drain into an idle evaporator. An alternative arrangement is shown in Figure 4D for cases where the compressor is above the evaporators. Figure 5 illustrates typical piping for evaporators above and below a common suction line. All horizontal runs should be level or pitched toward the compressor to ensure oil return. Traps shown in the suction lines after the evaporator suction outlet are recommended by thermal expansion valve manufacturers to

Halocarbon Refrigeration Systems

1.19

Fig. 4

Suction Line Piping at Evaporator Coils

Table 19 Minimum Refrigeration Capacity in Kilowatts for Oil Entrainment up Suction Risers (Copper Tubing, ASTM B88M Type B, Metric Size) Refrigerant 22

Saturated Suction Temp., Gas °C Temp., °C –40

–20

–5

5

134a

–10

–5

5

10

Tubing Nominal OD, mm 12

15

18

22

28

35

42

54

67

79

105

130

–35

0.182

0.334

0.561

0.956

1.817

3.223

5.203

9.977

14.258

26.155

53.963

93.419

–25

0.173

0.317

0.532

0.907

1.723

3.057

4.936

9.464

16.371

24.811

51.189

88.617

–15

0.168

0.307

0.516

0.880

1.672

2.967

4.791

9.185

15.888

24.080

49.681

86.006

–15

0.287

0.527

0.885

1.508

2.867

5.087

8.213

15.748

27.239

41.283

85.173 147.449

–5

0.273

0.501

0.841

1.433

2.724

4.834

7.804

14.963

25.882

39.226

80.929 140.102

5

0.264

0.485

0.815

1.388

2.638

4.680

7.555

14.487

25.058

37.977

78.353 135.642

0

0.389

0.713

1.198

2.041

3.879

6.883

11.112

21.306

36.854

55.856

115.240 199.499

10

0.369

0.676

1.136

1.935

3.678

6.526

10.535

20.200

34.940

52.954

109.254 189.136

20

0.354

0.650

1.092

1.861

3.537

6.275

10.131

19.425

33.600

50.924

105.065 181.884

10

0.470

0.862

1.449

2.468

4.692

8.325

13.441

25.771

44.577

67.560

139.387 241.302

20

0.440

0.807

1.356

2.311

4.393

7.794

12.582

24.126

41.731

63.246

130.488 225.896

30

0.422

0.774

1.301

2.217

4.213

7.476

12.069

23.141

40.027

60.665

125.161 216.675

–5

0.274

0.502

0.844

1.437

2.732

4.848

7.826

15.006

25.957

39.340

81.164 140.509

5

0.245

0.450

0.756

1.287

2.447

4.342

7.010

13.440

23.248

35.235

72.695 125.847

15

0.238

0.436

0.732

1.247

2.370

4.206

6.790

13.019

22.519

34.129

70.414 121.898

0

0.296

0.543

0.913

1.555

2.956

5.244

8.467

16.234

28.081

42.559

87.806 152.006

10

0.273

0.500

0.840

1.431

2.720

4.827

7.792

14.941

25.843

39.168

80.809 139.894

20

0.264

0.484

0.813

1.386

2.634

4.674

7.546

14.468

25.026

37.929

78.254 135.471

10

0.357

0.655

1.100

1.874

3.562

6.321

10.204

19.565

33.843

51.292

105.823 183.197

20

0.335

0.615

1.033

1.761

3.347

5.938

9.586

18.380

31.792

48.184

99.412 172.098

30

0.317

0.582

0.978

1.667

3.168

5.621

9.075

17.401

30.099

45.617

94.115 162.929

15

0.393

0.721

1.211

2.063

3.921

6.957

11.232

21.535

37.250

56.456

116.479 201.643

25

0.370

0.679

1.141

1.944

3.695

6.555

10.583

20.291

35.098

53.195

109.749 189.993

35

0.358

0.657

1.104

1.881

3.576

6.345

10.243

19.640

33.971

51.486

106.224 183.891

Refrigerant

20

30

50

1.17 1.20

1.08 1.10

0.91 0.89

Notes: 1. Refrigeration capacity in kilowatts is based on saturated evaporator as shown in table and condensing temperature of 40°C. For other liquid line temperatures, use correction factors in table at right. 2. Values computed using ISO 32 mineral oil for R-22 and R-502. R-134a computed using ISO 32 ester-based oil.

22 134a

Liquid Temperature, °C

1.20

Fig. 5

2014 ASHRAE Handbook—Refrigeration (SI)

Typical Piping from Evaporators Located above and below Common Suction Line

prevent erratic operation of the thermal expansion valve. Expansion valve bulbs are located on the suction lines between the evaporator and these traps. The traps serve as drains and help prevent liquid from accumulating under the expansion valve bulbs during compressor off cycles. They are useful only where straight runs or risers are encountered in the suction line leaving the evaporator outlet.

PIPING AT MULTIPLE COMPRESSORS Multiple compressors operating in parallel must be carefully piped to ensure proper operation.

Suction Piping Suction piping should be designed so that all compressors run at the same suction pressure and oil is returned in equal proportions. All suction lines should be brought into a common suction header to return oil to each crankcase as uniformly as possible. Depending on the type and size of compressors, oil may be returned by designing the piping in one or more of the following schemes: • Oil returned with the suction gas to each compressor • Oil contained with a suction trap (accumulator) and returned to the compressors through a controlled means • Oil trapped in a discharge line separator and returned to the compressors through a controlled means (see the section on Discharge Piping) The suction header is a means of distributing suction gas equally to each compressor. Header design can freely pass the suction gas and oil mixture or provide a suction trap for the oil. The header should be run above the level of the compressor suction inlets so oil can drain into the compressors by gravity. Figure 6 shows a pyramidal or yoke-type suction header to maximize pressure and flow equalization at each of three compressor suction inlets piped in parallel. This type of construction is recommended for applications of three or more compressors in parallel. For two compressors in parallel, a single feed between the two compressor takeoffs is acceptable. Although not as good for equalizing flow and pressure drops to all compressors, one alternative is to have the suction line from evaporators enter at one end of the header instead of using the yoke arrangement. The suction header may have to be enlarged to minimize pressure drop and flow turbulence. Suction headers designed to freely pass the gas/oil mixture should have branch suction lines to compressors connected to the side of the header. Return mains from the evaporators should not be connected into the suction header to form crosses with the branch suction lines to the compressors. The header should be full size based on the largest mass flow of the suction line returning to the

Fig. 6 Suction and Hot-Gas Headers for Multiple Compressors compressors. Takeoffs to the compressors should either be the same size as the suction header or be constructed so that oil will not trap in the suction header. Branch suction lines to the compressors should not be reduced until the vertical drop is reached. Suction traps are recommended wherever (1) parallel compressors, (2) flooded evaporators, (3) double suction risers, (4) long suction lines, (5) multiple expansion valves, (6) hot-gas defrost, (7) reverse-cycle operation, or (8) suction-pressure regulators are used. Depending on system size, the suction header may be designed to function as a suction trap. The suction header should be large enough to provide a low-velocity region in the header to allow suction gas and oil to separate. See the section on Low-Pressure Receiver Sizing in Chapter 4 to find recommended velocities for separation. Suction gas flow for individual compressors should be taken off the top of the suction header. Oil can be returned to the compressor directly or through a vessel equipped with a heater to boil off refrigerant and then allow oil to drain to the compressors or other devices used to feed oil to the compressors. The suction trap must be sized for effective gas and liquid separation. Adequate liquid volume and a means of disposing of it must be provided. A liquid transfer pump or heater may be used. Chapter 4 has further information on separation and liquid transfer pumps. An oil receiver equipped with a heater effectively evaporates liquid refrigerant accumulated in the suction trap. It also ensures that each compressor receives its share of oil. Either crankcase float valves or external float switches and solenoid valves can be used to control the oil flow to each compressor. A gravity-feed oil receiver should be elevated to overcome the pressure drop between it and the crankcase. The oil receiver should be sized so that a malfunction of the oil control mechanism cannot overfill an idle compressor. Figure 7 shows a recommended hookup of multiple compressors, suction trap (accumulator), oil receiver, and discharge line oil separators. The oil receiver also provides a reserve supply of oil for compressors where oil in the system outside the compressor varies with system loading. The heater mechanism should always be submerged.

Discharge Piping The piping arrangement in Figure 6 is suggested for discharge piping. The piping must be arranged to prevent refrigerant liquid and oil from draining back into the heads of idle compressors. A check valve in the discharge line may be necessary to prevent refrigerant and oil from entering the compressor heads by migration. It is recommended that, after leaving the compressor head, the piping be routed to a lower elevation so that a trap is formed to allow drainback of refrigerant and oil from the discharge line when flow rates

Halocarbon Refrigeration Systems

1.21

Fig. 8 Interconnecting Piping for Multiple Condensing Units Fig. 7 Parallel Compressors with Gravity Oil Flow are reduced or the compressors are off. If an oil separator is used in the discharge line, it may suffice as the trap for drainback for the discharge line. Avoid using a bullheaded tee at the junction of two compressor branches and the main discharge header: this configuration causes increased turbulence, increased pressure drop, and possible hammering in the line. When an oil separator is used on multiple-compressor arrangements, oil must be piped to return to the compressors. This can be done in various ways, depending on the oil management system design. Oil may be returned to an oil receiver that is the supply for control devices feeding oil back to the compressors.

Interconnecting Crankcases When two or more compressors are interconnected, a method must be provided to equalize the crankcases. Some compressor designs do not operate correctly with simple equalization of the crankcases. For these systems, it may be necessary to design a positive oil float control system for each compressor crankcase. A typical system allows oil to collect in a receiver that, in turn, supplies oil to a device that meters it back into the compressor crankcase to maintain a proper oil level (Figure 7). Compressor systems that can be equalized should be placed on foundations so that all oil equalizer tapping locations are exactly level. If crankcase floats (as in Figure 7) are not used, an oil equalization line should connect all crankcases to maintain uniform oil levels. The oil equalizer may be run level with the tapping, or, for convenient access to compressors, it may be run at the floor (Figure 8). It should never be run at a level higher than that of the tapping. For the oil equalizer line to work properly, equalize the crankcase pressures by installing a gas equalizer line above the oil level. This line may be run to provide head room (Figure 8) or run level with tapping on the compressors. It should be piped so that oil or liquid refrigerant will not be trapped. Both lines should be the same size as the tapping on the largest compressor and should be valved so that any one machine can be taken out for repair. The piping should be arranged to absorb vibration.

PIPING AT VARIOUS SYSTEM COMPONENTS Flooded Fluid Coolers For a description of flooded fluid coolers, see Chapter 42 of the 2012 ASHRAE Handbook—HVAC Systems and Equipment.

Fig. 9 Typical Piping at Flooded Fluid Cooler Shell-and-tube flooded coolers designed to minimize liquid entrainment in the suction gas require a continuous liquid bleed line (Figure 9) installed at some point in the cooler shell below the liquid level to remove trapped oil. This continuous bleed of refrigerant liquid and oil prevents the oil concentration in the cooler from getting too high. The location of the liquid bleed connection on the shell depends on the refrigerant and oil used. For refrigerants that are highly miscible with the oil, the connection can be anywhere below the liquid level. Refrigerant 22 can have a separate oil-rich phase floating on a refrigerant-rich layer. This becomes more pronounced as evaporating temperature drops. When R-22 is used with mineral oil, the bleed line is usually taken off the shell just slightly below the liquid level, or there may be more than one valved bleed connection at slightly different levels so that the optimum point can be selected during operation. With alkyl benzene lubricants, oil/refrigerant miscibility may be high enough that the oil bleed connection can be anywhere below the liquid level. The solubility charts in Chapter 12 give specific information. Where the flooded cooler design requires an external surge drum to separate liquid carryover from suction gas off the tube bundle, the richest oil concentration may or may not be in the cooler. In some cases, the surge drum has the highest concentration of oil. Here, the refrigerant and oil bleed connection is taken from the surge drum. The refrigerant and oil bleed from the cooler by gravity. The bleed sometimes drains into the suction line so oil can be returned to the

1.22 compressor with the suction gas after the accompanying liquid refrigerant is vaporized in a liquid-suction heat interchanger. A better method is to drain the refrigerant/oil bleed into a heated receiver that boils refrigerant off to the suction line and drains oil back to the compressor.

Refrigerant Feed Devices For further information on refrigerant feed devices, see Chapter 11. The pilot-operated low-side float control (Figure 9) is sometimes selected for flooded systems using halocarbon refrigerants. Except for small capacities, direct-acting low-side float valves are impractical for these refrigerants. The displacer float controlling a pneumatic valve works well for low-side liquid level control; it allows the cooler level to be adjusted within the instrument without disturbing the piping. High-side float valves are practical only in single-evaporator systems, because distribution problems result when multiple evaporators are used. Float chambers should be located as near the liquid connection on the cooler as possible because a long length of liquid line, even if insulated, can pick up room heat and give an artificial liquid level in the float chamber. Equalizer lines to the float chamber must be amply sized to minimize the effect of heat transmission. The float chamber and its equalizing lines must be insulated. Each flooded cooler system must have a way of keeping oil concentration in the evaporator low, both to minimize the bleedoff needed to keep oil concentration in the cooler low and to reduce system losses from large stills. A highly efficient discharge gas/oil separator can be used for this purpose. At low temperatures, periodic warm-up of the evaporator allows recovery of oil accumulation in the chiller. If continuous operation is required, dual chillers may be needed to deoil an oil-laden evaporator, or an oil-free compressor may be used.

2014 ASHRAE Handbook—Refrigeration (SI) consequently, no heat is available from the high-pressure liquid, and the cooler must starve itself to obtain the superheat necessary to open the valve. When the valve does open, excessive superheat causes it to overfeed until the bulb senses liquid downstream from the interchanger. Therefore, the remote bulb should be positioned between the cooler and the interchanger. Figure 11 shows a typical piping arrangement that has been successful in packaged water chillers with DX coolers. With this arrangement, automatic recycling pumpdown is needed on the lag compressor to prevent leakage through compressor valves, allowing migration to the cold evaporator circuit. It also prevents liquid from slugging the compressor at start-up. On larger systems, the limited size of thermostatic expansion valves may require use of a pilot-operated liquid valve controlled by a small thermostatic expansion valve (Figure 12). The equalizing connection and bulb of the pilot thermostatic expansion valve should be treated as a direct-acting thermal expansion valve. A small solenoid valve in the pilot line shuts off the high side from the low during shutdown. However, the main liquid valve does not open and close instantaneously.

Direct-Expansion Air Coils For further information on these coils, see Chapter 23 of the 2012 ASHRAE Handbook—HVAC Systems and Equipment. The most common ways of arranging DX coils are shown in Figures 13 and 14. The method shown in Figure 14 provides the superheat needed to operate the thermostatic expansion valve and is effective for heat transfer because leaving air contacts the coldest evaporator surface. This arrangement is advantageous on low-temperature applications,

Direct-Expansion Fluid Chillers For details on these chillers, see Chapter 43 in the 2012 ASHRAE Handbook—HVAC Systems and Equipment. Figure 10 shows typical piping connections for a multicircuit direct-expansion (DX) chiller. Each circuit contains its own thermostatic expansion and solenoid valves. One solenoid valve can be wired to close at reduced system capacity. The thermostatic expansion valve bulbs should be located between the cooler and the liquid-suction interchanger, if used. Locating the bulb downstream from the interchanger can cause excessive cycling of the thermostatic expansion valve because the flow of high-pressure liquid through the interchanger ceases when the thermostatic expansion valve closes; Fig. 11 Typical Refrigerant Piping in Liquid Chilling Package with Two Completely Separate Circuits

Fig. 10 Two-Circuit Direct-Expansion Cooler Connections (for Single-Compressor System)

Fig. 12 Direct-Expansion Cooler with Pilot-Operated Control Valve

Halocarbon Refrigeration Systems where the coil pressure drop represents an appreciable change in evaporating temperature. Direct-expansion air coils can be located in any position as long as proper refrigerant distribution and continuous oil removal facilities are provided. Figure 13 shows top-feed, free-draining piping with a vertical up-airflow coil. In Figure 14, which illustrates a horizontal-airflow coil, suction is taken off the bottom header connection, providing free oil draining. Many coils are supplied with connections at each end of the suction header so that a free-draining connection can be used regardless of which side of the coil is up; the other end is then capped. In Figure 15, a refrigerant upfeed coil is used with a vertical downflow air arrangement. Here, the coil design must provide sufficient gas velocity to entrain oil at lowest loadings and to carry it into the suction line.

1.23 Pumpdown compressor control is desirable on all systems using downfeed or upfeed evaporators, to protect the compressor against a liquid slugback in cases where liquid can accumulate in the suction header and/or the coil on system off cycles. Pumpdown compressor control is described in the section on Keeping Liquid from Crankcase During Off Cycles. Thermostatic expansion valve operation and application are described in Chapter 11. Thermostatic expansion valves should be sized carefully to avoid undersizing at full load and oversizing at partial load. The refrigerant pressure drops through the system (distributor, coil, condenser, and refrigerant lines, including liquid lifts) must be properly evaluated to determine the correct pressure drop available across the valve on which to base the selection. Variations in condensing pressure greatly affect the pressure available across the valve, and hence its capacity. Oversized thermostatic expansion valves result in cycling that alternates flooding and starving the coil. This occurs because the valve attempts to throttle at a capacity below its capability, which causes periodic flooding of the liquid back to the compressor and wide temperature variations in the air leaving the coil. Reduced compressor capacity further aggravates this problem. Systems having multiple coils can use solenoid valves located in the liquid line feeding each evaporator or group of evaporators to close them off individually as compressor capacity is reduced. For information on defrosting, see Chapter 14.

Flooded Evaporators

Fig. 13 Direct-Expansion Evaporator (Top-Feed, Free-Draining)

Fig. 14

Direct-Expansion Evaporator (Horizontal Airflow)

Flooded evaporators may be desirable when a small temperature differential is required between the refrigerant and the medium being cooled. A small temperature differential is advantageous in low-temperature applications. In a flooded evaporator, the coil is kept full of refrigerant when cooling is required. The refrigerant level is generally controlled through a high- or low-side float control. Figure 16 represents a typical arrangement showing a low-side float control, oil return line, and heat interchanger. Circulation of refrigerant through the evaporator depends on gravity and a thermosiphon effect. A mixture of liquid refrigerant and vapor returns to the surge tank, and the vapor flows into the suction line. A baffle installed in the surge tank helps prevent foam and liquid from entering the suction line. A liquid refrigerant circulating pump (Figure 17) provides a more positive way of obtaining a high circulation rate. Taking the suction line off the top of the surge tank causes difficulties if no special provisions are made for oil return. For this reason, the oil return lines in Figure 16 should be installed. These lines are connected near the bottom of the float chamber and also just

Fig. 15 Direct-Expansion Evaporator (Bottom-Feed)

1.24

2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 18 Double Hot-Gas Riser

Fig. 16 Flooded Evaporator (Gravity Circulation)

Fig. 17

Flooded Evaporator (Forced Circulation)

below the liquid level in the surge tank (where an oil-rich liquid refrigerant exists). They extend to a lower point on the suction line to allow gravity flow. Included in this oil return line is (1) a solenoid valve that is open only while the compressor is running and (2) a metering valve that is adjusted to allow a constant but small-volume return to the suction line. A liquid-line sight glass may be installed downstream from the metering valve to serve as a convenient check on liquid being returned. Oil can be returned satisfactorily by taking a bleed of refrigerant and oil from the pump discharge (Figure 17) and feeding it to the heated oil receiver. If a low-side float is used, a jet ejector can be used to remove oil from the quiescent float chamber.

DISCHARGE (HOT-GAS) LINES Hot-gas lines should be designed to • Avoid trapping oil at part-load operation • Prevent condensed refrigerant and oil in the line from draining back to the head of the compressor • Have carefully selected connections from a common line to multiple compressors • Avoid developing excessive noise or vibration from hot-gas pulsations, compressor vibration, or both

Oil Transport up Risers at Normal Loads. Although a low pressure drop is desired, oversized hot-gas lines can reduce gas velocities to a point where the refrigerant will not transport oil. Therefore, when using multiple compressors with capacity control, hot-gas risers must transport oil at all possible loadings. Minimum Gas Velocities for Oil Transport in Risers. Minimum capacities for oil entrainment in hot-gas line risers are shown in Table 20. On multiple-compressor installations, the lowest possible system loading should be calculated and a riser size selected to give at least the minimum capacity indicated in the table for successful oil transport. In some installations with multiple compressors and with capacity control, a vertical hot-gas line, sized to transport oil at minimum load, has excessive pressure drop at maximum load. When this problem exists, either a double riser or a single riser with an oil separator can be used. Double Hot-Gas Risers. A double hot-gas riser can be used the same way it is used in a suction line. Figure 18 shows the double riser principle applied to a hot-gas line. Its operating principle and sizing technique are described in the section on Double Suction Risers. Single Riser and Oil Separator. As an alternative, an oil separator in the discharge line just before the riser allows sizing the riser for a low pressure drop. Any oil draining back down the riser accumulates in the oil separator. With large multiple compressors, separator capacity may dictate the use of individual units for each compressor located between the discharge line and the main discharge header. Horizontal lines should be level or pitched downward in the direction of gas flow to facilitate travel of oil through the system and back to the compressor. Piping to Prevent Liquid and Oil from Draining to Compressor Head. Whenever the condenser is located above the compressor, the hot-gas line should be trapped near the compressor before rising to the condenser, especially if the hot-gas riser is long. This minimizes the possibility of refrigerant, condensed in the line during off cycles, draining back to the head of the compressor. Also, any oil traveling up the pipe wall will not drain back to the compressor head. The loop in the hot-gas line (Figure 19) serves as a reservoir and traps liquid resulting from condensation in the line during shutdown, thus preventing gravity drainage of liquid and oil back to the compressor head. A small high-pressure float drainer should be installed at the bottom of the trap to drain any significant amount of refrigerant condensate to a low-side component such as a suction accumulator or low-pressure receiver. This float prevents excessive build-up of liquid in the trap and possible liquid hammer when the compressor is restarted. For multiple-compressor arrangements, each discharge line should have a check valve to prevent gas from active compressors from condensing on heads of idle compressors. For single-compressor applications, a tightly closing check valve should be installed in the hot-gas line of the compressor whenever

Halocarbon Refrigeration Systems Table 20

Minimum Refrigeration Capacity in Kilowatts for Oil Entrainment up Hot-Gas Risers (Copper Tubing, ASTM B88M Type B, Metric Size)

Saturated Discharge Discharge Gas Temp., Temp., Refrigerant °C °C 22

134a

1.25

Tubing Diameter, Nominal OD, mm 12

15

18

22

28

35

42

54

67

20

60 70 80

0.563 0.549 0.535

0.032 1.006 0.982

0.735 1.691 1.650

2.956 2.881 2.811

5.619 5.477 5.343

9.969 9.717 9.480

16.094 15.687 15.305

30.859 30.078 29.346

43.377 52.027 50.761

80.897 116.904 288.938 48.851 162.682 281.630 76.933 158.726 173.780

30

70 80 90

0.596 0.579 0.565

1.092 1.062 0.035

1.836 1.785 1.740

3.127 3.040 2.964

5.945 5.779 5.635

10.547 10.254 9.998

17.028 16.554 16.140

32.649 31.740 30.948

56.474 54.901 53.531

85.591 176.588 305.702 83.208 171.671 297.190 81.131 167.386 289.773

40

80 90 100

0.618 0.601 0.584

1.132 1.103 1.071

1.903 1.853 1.800

3.242 3.157 3.067

6.163 6.001 5.830

10.934 10.647 10.343

17.653 17.189 16.698

33.847 32.959 32.018

58.546 47.009 55.382

88.732 183.069 316.922 86.403 178.263 308.603 83.936 173.173 299.791

50

90 100 110

0.630 0.611 0.595

1.156 1.121 1.092

1.943 1.884 1.834

3.310 3.209 3.125

6.291 6.100 5.941

11.162 10.823 10.540

18.020 17.473 17.016

34.552 33.503 32.627

59.766 57.951 56.435

90.580 186.882 323.523 87.831 181.209 313.702 85.532 176.467 305.493

20

60 70 80

0.469 0.441 0.431

0.860 0.808 0.790

1.445 1.358 1.327

2.462 2.314 2.261

4.681 4.399 4.298

8.305 7.805 7.626

13.408 12.600 12.311

25.709 24.159 23.605

44.469 41.788 40.830

67.396 139.050 240.718 63.334 130.668 226.207 61.881 127.671 221.020

30

70 80 90

0.493 0.463 0.452

0.904 0.849 0.829

1.519 1.426 1.393

2.587 2.430 2.374

4.918 4.260 4.513

8.726 8.196 8.007

14.087 13.232 12.926

27.011 25.371 24.785

46.722 43.885 42.870

70.812 145.096 252.916 66.512 137.225 237.560 64.974 134.052 232.066

40

80 90 100

0.507 0.477 0.465

0.930 0.874 0.852

1.563 1.469 1.432

2.662 2.502 2.439

5.061 4.756 4.637

8.979 8.439 8.227

14.496 13.624 13.281

27.794 26.122 25.466

48.075 45.184 44.048

72.863 150.328 260.242 68.480 141.285 244.588 66.759 137.735 238.443

50

90 100 110

0.510 0.479 0.467

0.936 0.878 0.857

1.573 1.476 1.441

2.679 2.514 2.454

5.093 4.779 4.665

9.037 8.480 8.278

14.589 13.690 13.364

27.973 26.248 25.624

48.385 45.402 44.322

73.332 151.296 261.918 68.811 141.969 245.772 67.173 138.590 239.921

–50 0.87 —

–40 0.90 —

Notes: 1. Refrigeration capacity in kilowatts is based on saturated evaporator at –5°C, and condensing temperature as shown in table. For other liquid line temperatures, use correction factors in table at right. 2. Values computed using ISO 32 mineral oil for R-22, and ISO 32 esterbased oil for R-134a.

Fig. 19 Hot-Gas Loop

79

105

130

Saturated Suction Temperature, °C Refrigerant 22 134a

–30 0.93 —

–20 0.96 —

0 — 1.02

5 1.02 1.04

10 — 1.06

the condenser and the receiver ambient temperature are higher than that of the compressor. The check valve prevents refrigerant from boiling off in the condenser or receiver and condensing on the compressor heads during off cycles. This check valve should be a piston type, which closes by gravity when the compressor stops running. A spring-loaded check may incur chatter (vibration), particularly on slow-speed reciprocating compressors. For compressors equipped with water-cooled oil coolers, a water solenoid and water-regulating valve should be installed in the water line so that the regulating valve maintains adequate cooling during operation, and the solenoid stops flow during the off cycle to prevent localized condensing of the refrigerant. Hot-Gas (Discharge) Mufflers. Mufflers can be installed in hot-gas lines to dampen discharge gas pulsations, reducing vibration and noise. Mufflers should be installed in a horizontal or downflow portion of the hot-gas line immediately after it leaves the compressor. Because gas velocity through the muffler is substantially lower than that through the hot-gas line, the muffler may form an oil trap. The muffler should be installed to allow oil to flow through it and not be trapped.

1.26

2014 ASHRAE Handbook—Refrigeration (SI) DEFROST GAS SUPPLY LINES

Sizing refrigeration lines to supply defrost gas to one or more evaporators is not an exact science. The parameters associated with sizing the defrost gas line are related to allowable pressure drop and refrigerant flow rate during defrost. Engineers use an estimated two times the evaporator load for effective refrigerant flow rate to determine line sizing requirements. Pressure drop is not as critical during the defrost cycle, and many engineers use velocity as the criterion for determining line size. The effective condensing temperature and average temperature of the gas must be determined. The velocity determined at saturated conditions gives a conservative line size. Controlled testing (Stoecker 1984) showed that, in small coils with R-22, the defrost flow rate tends to be higher as the condensing temperature increases. The flow rate is on the order of two to three times the normal evaporator flow rate, which supports the estimated two times used by practicing engineers.

check valve in the vent with flow in the direction of the condenser. The check valve should be selected for minimum opening pressure (i.e., approximately 3.5 kPa). When determining condensate drop leg height, allowance must be made to overcome both the pressure drop across this check valve and the refrigerant pressure drop through the condenser. This ensures that there will be no liquid

HEAT EXCHANGERS AND VESSELS Receivers Refrigerant receivers are vessels used to store excess refrigerant circulated throughout the system. Their purpose is to • Provide pumpdown storage capacity when another part of the system must be serviced or the system must be shut down for an extended time. In some water-cooled condenser systems, the condenser also serves as a receiver if the total refrigerant charge does not exceed its storage capacity. • Handle the excess refrigerant charge needed by air-cooled condensers that require flooding to maintain minimum condensing pressures (see the section on Pressure Control for Refrigerant Condensers). • Receive refrigerant draining from the condenser, to allow the condenser to maintain its usable surface area for condensing. • Accommodate a fluctuating charge in the low side on systems where the operating charge in the evaporator varies for different loading conditions. When an evaporator is fed with a thermal expansion valve, hand expansion valve, or low-pressure float, the operating charge in the evaporator varies considerably depending on the loading. During low load, the evaporator requires a larger charge because boiling is not as intense. When load increases, the operating charge in the evaporator decreases, and the receiver must store excess refrigerant. Connections for Through-Type Receiver. When a throughtype receiver is used, liquid must always flow from condenser to receiver. Pressure in the receiver must be lower than that in the condenser outlet. The receiver and its associated piping provide free flow of liquid from the condenser to the receiver by equalizing pressures between the two so that the receiver cannot build up a higher pressure than the condenser. If a vent is not used, piping between condenser and receiver (condensate line) is sized so that liquid flows in one direction and gas flows in the opposite direction. Sizing the condensate line for 0.5 m/s liquid velocity is usually adequate to attain this flow. Piping should slope at least 20 mm/m and eliminate any natural liquid traps. Figure 20 illustrates this configuration. Piping between the condenser and the receiver can be equipped with a separate vent (equalizer) line to allow receiver and condenser pressures to equalize. This external vent line can be piped either with or without a check valve in the vent line (see Figures 22 and 23). If there is no check valve, prevent discharge gas from discharging directly into the vent line; this should prevent a gas velocity pressure component from being introduced on top of the liquid in the receiver. When the piping configuration is unknown, install a

Fig. 20 Shell-and-Tube Condenser to Receiver Piping (Through-Type Receiver)

Fig. 21

Shell-and-Tube Condenser to Receiver Piping (Surge-Type Receiver)

Fig. 22 Parallel Condensers with Through-Type Receiver

Pipe Size Copper Nominal mm

R-22 Mass Flow Data, kg/s

R-134a Mass Flow Data, kg/s R-404a Mass Flow Data, kg/s

Velocity, m/s 5

Refrigerant Flow Capacity Data For Defrost Lines

Velocity, m/s 5

R-507 Mass Flow Data, kg/s

Velocity, m/s 5

R-410a Mass Flow Data, kg/s R-407c Mass Flow Data, kg/s

Velocity, m/s 5

Velocity, m/s 5

Velocity, m/s

10

15

10

15

10

15

10

15

10

15

10

15

12

0.012

0.024

0.035

0.016

0.032

0.049

0.024

0.047

0.071

0.025

0.050

0.075

0.024

0.048

0.071

0.016

5

0.032

0.048

15

0.019

0.038

0.057

0.026

0.053

0.079

0.039

0.077

0.116

0.041

0.081

0.122

0.039

0.077

0.116

0.026

0.051

0.077

18

0.029

0.058

0.087

0.040

0.080

0.119

0.058

0.117

0.175

0.062

0.123

0.185

0.059

0.117

0.176

0.039

0.078

0.117

22

0.044

0.088

0.133

0.061

0.122

0.183

0.089

0.179

0.268

0.094

0.189

0.283

0.090

0.179

0.269

0.060

0.119

0.179

28

0.074

0.148

0.222

0.102

0.204

0.305

0.149

0.299

0.448

0.158

0.316

0.474

0.150

0.300

0.450

0.100

0.200

0.299

35

0.120

0.230

0.350

0.160

0.320

0.480

0.236

0.473

0.709

0.250

0.500

0.750

0.237

0.474

0.711

0.158

0.316

0.474

42

0.170

0.340

0.510

0.240

0.470

0.710

0.347

0.694

1.041

0.367

0.733

1.100

0.348

0.696

1.044

0.232

0.463

0.695

54

0.290

0.580

0.870

0.400

0.800

1.190

0.584

1.168

1.752

0.617

1.234

1.851

0.586

1.171

1.757

0.390

0.780

1.169

67

0.450

0.890

1.340

0.620

1.230

1.850

0.905

1.811

2.716

0.956

1.913

2.869

0.908

1.816

2.723

0.604

1.208

1.813

79

0.620

1.250

1.870

0.860

1.720

2.580

1.263

2.525

3.788

1.334

2.668

4.002

1.266

2.532

3.798

0.843

1.685

2.528

105

1.110

2.230

3.340

1.530

3.070

4.600

2.254

4.507

6.761

2.381

4.762

7.143

2.260

4.520

6.780

1.504

3.008

4.512

130

1.730

3.460

5.180

2.380

4.760

7.140

3.496

6.992

10.488

3.693

7.387

11.080

3.505

7.011

10.516

2.333

4.666

6.999

156

2.500

5.010

7.510

3.450

6.900

10.300

5.064

10.128

15.192

5.350

10.700

16.050

5.078

10.156

15.233

3.380

6.759

10.139

206

4.330

8.660

13.000

5.970

11.900

17.900

8.762

17.525

26.287

9.258

18.516

27.773

8.787

17.573

26.360

5.848

11.696

17.544

257

6.730

13.500

20.200

9.280

18.600

27.800

13.624

27.248

40.873

14.395

28.789

43.184

13.662

27.324

40.985

9.093

18.186

27.279

10

0.018

0.035

0.053

0.024

0.049

0.073

0.026

0.053

0.079

0.028

0.056

0.083

0.026

0.053

0.079

0.018

0.035

0.053

15

0.028

0.056

0.084

0.039

0.078

0.116

0.044

0.088

0.132

0.046

0.093

0.139

0.044

0.088

0.132

0.029

0.059

0.088

20

0.049

0.099

0.148

0.068

0.136

0.204

0.081

0.162

0.243

0.086

0.171

0.257

0.081

0.162

0.244

0.054

0.108

0.162

25

0.080

0.160

0.240

0.110

0.220

0.330

0.135

0.270

0.404

0.142

0.285

0.427

0.135

0.270

0.405

0.090

0.180

0.270

32

0.139

0.280

0.420

0.191

0.382

0.570

0.240

0.481

0.721

0.254

0.508

0.762

0.241

0.482

0.723

0.160

0.321

0.481

40

0.190

0.380

0.570

0.260

0.520

0.780

0.331

0.662

0.993

0.350

0.700

1.049

0.332

0.664

0.996

0.221

0.442

0.663

50

0.310

0.620

0.930

0.430

0.860

1.280

0.629

1.257

1.886

0.664

1.329

1.993

0.630

1.261

1.891

0.420

0.839

1.259

65

0.440

0.890

1.330

0.610

1.220

1.830

0.896

1.793

2.689

0.947

1.894

2.841

0.899

1.798

2.696

0.598

1.196

1.795

Halocarbon Refrigeration Systems

Table 21

Steel Nominal mm

0.680

1.370

2.050

0.940

1.890

2.830

1.384

2.768

4.153

1.462

2.925

4.387

1.388

2.776

4.164

0.924

1.848

2.771

1.180

2.360

3.540

1.620

3.250

4.870

2.385

4.770

7.156

2.520

5.040

7.560

2.392

4.784

7.175

1.592

3.184

4.776

125

1.850

3.700

5.550

2.550

5.100

7.650

3.745

7.491

11.236

3.957

7.914

11.871

3.756

7.511

11.267

2.500

4.999

7.499

150

2.680

5.350

8.030

3.690

7.370

11.100

5.413

10.826

16.239

5.719

11.438

17.157

5.428

10.856

16.284

3.613

7.225

10.838

200

4.630

9.260

13.900

6.380

12.800

19.100

9.373

18.747

28.120

9.903

19.806

29.710

9.399

18.798

28.197

6.256

12.512

18.767

250

7.300

14.600

21.900

10.100

20.100

30.200

14.774

29.549

44.323

15.610

31.220

46.829

14.815

29.630

44.446

9.861

19.721

29.582

300

10.500

20.900

31.400

14.400

28.900

43.300

21.190

42.381

63.571

22.388

44.777

67.165

21.249

42.498

63.747

14.143

28.285

42.428

350

—

—

—

—

—

—

25.835

51.670

77.505

27.296

54.591

81.887

25.906

51.813

77.719

17.242

34.485

51.727

400

—

—

—

—

—

—

34.223

68.446

102.669

36.158

72.315

108.473

34.317

68.635

102.952

22.840

45.681

68.521

Note: Refrigerant flow data based on saturated condensing temperature of 21°C.

1.27

80 100

1.28 back-up into an operating condenser on a multiple-condenser application when one or more of the condensers is idle. The condensate line should be sized so that velocity does not exceed 0.75 m/s. The vent line flow is from receiver to condenser when receiver temperature is higher than condensing temperature. Flow is from condenser to receiver when air temperature around the receiver is below condensing temperature. Flow rate depends on this temperature difference as well as on the receiver surface area. Vent size can be calculated from this flow rate. Connections for Surge-Type Receiver. The purpose of a surgetype receiver is to allow liquid to flow to the expansion valve without exposure to refrigerant in the receiver, so that it can remain subcooled. The receiver volume is available for liquid that is to be removed from the system. Figure 21 shows an example of connections for a surgetype receiver. Height h must be adequate for a liquid pressure at least as large as the pressure loss through the condenser, liquid line, and vent line at the maximum temperature difference between the receiver ambient and the condensing temperature. Condenser pressure drop at the greatest expected heat rejection should be obtained from the manufacturer. The minimum value of h can then be calculated to determine whether the available height will allow the surgetype receiver. Multiple Condensers. Two or more condensers connected in series or in parallel can be used in a single refrigeration system. If connected in series, the pressure losses through each condenser must be added. Condensers are more often arranged in parallel. Pressure loss through any one of the parallel circuits is always equal to that through any of the others, even if it results in filling much of one circuit with liquid while gas passes through another. Figure 22 shows a basic arrangement for parallel condensers with a through-type receiver. Condensate drop legs must be long enough to allow liquid levels in them to adjust to equalize pressure losses between condensers at all operating conditions. Drop legs should be 150 to 300 mm higher than calculated to ensure that liquid outlets drain freely. This height provides a liquid pressure to offset the largest condenser pressure loss. The liquid seal prevents gas blow-by between condensers. Large single condensers with multiple coil circuits should be piped as though the independent circuits were parallel condensers. For example, if the left condenser in Figure 22 has 14 kPa more pressure drop than the right condenser, the liquid level on the left is about 1.2 m higher than that on the right. If the condensate lines do not have enough vertical height for this level difference, liquid will back up into the condenser until pressure drop is the same through both circuits. Enough surface may be covered to reduce condenser capacity significantly. Condensate drop legs should be sized based on 0.75 m/s velocity. The main condensate lines should be based on 0.5 m/s. Depending on prevailing local and/or national safety codes, a relief device may have to be installed in the discharge piping. Figure 23 shows a piping arrangement for parallel condensers with a surge-type receiver. When the system is operating at reduced load, flow paths through the circuits may not be symmetrical. Small pressure differences are not unusual; therefore, the liquid line junction should be about 600 to 900 mm below the bottom of the condensers. The exact amount can be calculated from pressure loss through each path at all possible operating conditions. When condensers are water-cooled, a single automatic water valve for the condensers in one refrigeration system should be used. Individual valves for each condenser in a single system cannot maintain the same pressure and corresponding pressure drops. With evaporative condensers (Figure 24), pressure loss may be high. If parallel condensers are alike and all are operated, the differences may be small, and condenser outlets need not be more than 600 to 900 mm above the liquid line junction. If fans on one condenser are not operated while the fans on another condenser are,

2014 ASHRAE Handbook—Refrigeration (SI) then the liquid level in the one condenser must be high enough to compensate for the pressure drop through the operating condenser. When the available level difference between condenser outlets and the liquid-line junction is sufficient, the receiver may be vented to the condenser inlets (Figure 25). In this case, the surge-type receiver can be used. The level difference must then be at least equal to the greatest loss through any condenser circuit plus the greatest vent line loss when the receiver ambient is greater than the condensing temperature.

Air-Cooled Condensers Refrigerant pressure drop through air-cooled condensers must be obtained from the supplier for the particular unit at the specified load. If refrigerant pressure drop is low enough and the arrangement is practical, parallel condensers can be connected to allow for capacity reduction to zero on one condenser without causing liquid back-up in active condensers (Figure 26). Multiple condensers with high pressure drops can be connected as shown in Figure 26, provided that (1) the ambient at the receiver is equal to or lower than the inlet air temperature to the condenser; (2) capacity control affects all units equally; (3) all units operate when one operates, unless valved off at both inlet and outlet; and (4) all units are of equal size.

Fig. 23 Parallel Condensers with Surge-Type Receiver

Fig. 24 Single-Circuit Evaporative Condenser with Receiver and Liquid Subcooling Coil

Halocarbon Refrigeration Systems A single condenser with any pressure drop can be connected to a receiver without an equalizer and without trapping height if the condenser outlet and the line from it to the receiver can be sized for sewer flow without a trap or restriction, using a maximum velocity of 0.5 m/s. A single condenser can also be connected with an equalizer line to the hot-gas inlet if the vertical drop leg is sufficient to balance refrigerant pressure drop through the condenser and liquid line to the receiver. If unit sizes are unequal, additional liquid height H, equivalent to the difference in full-load pressure drop, is required. Usually, condensers of equal size are used in parallel applications. If the receiver cannot be located in an ambient temperature below the inlet air temperature for all operating conditions, sufficient extra height of drop leg H is required to overcome the

1.29 equivalent differences in saturation pressure of the receiver and the condenser. Subcooling by the liquid leg tends to condense vapor in the receiver to reach a balance between rate of condensation, at an intermediate saturation pressure, and heat gain from ambient to the receiver. A relatively large liquid leg is required to balance a small temperature difference; therefore, this method is probably limited to marginal cases. Liquid leaving the receiver is nonetheless saturated, and any subcooling to prevent flashing in the liquid line must be obtained downstream of the receiver. If the temperature of the receiver ambient is above the condensing pressure only at part-load conditions, it may be acceptable to back liquid into the condensing surface, sacrificing the operating economy of lower part-load pressure for a lower liquid leg requirement. The receiver must be adequately sized to contain a minimum of the backed-up liquid so that the condenser can be fully drained when full load is required. If a low-ambient control system of backing liquid into the condenser is used, consult the system supplier for proper piping.

REFRIGERATION ACCESSORIES Liquid-Suction Heat Exchangers Generally, liquid-suction heat exchangers subcool liquid refrigerant and superheat suction gas. They are used for one or more of the following functions:

Fig. 25 Multiple Evaporative Condensers with Equalization to Condenser Inlets

Fig. 26 Multiple Air-Cooled Condensers

• Increasing efficiency of the refrigeration cycle. Efficiency of the thermodynamic cycle of certain halocarbon refrigerants can be increased when the suction gas is superheated by removing heat from the liquid. This increased efficiency must be evaluated against the effect of pressure drop through the suction side of the exchanger, which forces the compressor to operate at a lower suction pressure. Liquid-suction heat exchangers are most beneficial at low suction temperatures. The increase in cycle efficiency for systems operating in the air-conditioning range (down to about –1°C evaporating temperature) usually does not justify their use. The heat exchanger can be located wherever convenient. • Subcooling liquid refrigerant to prevent flash gas at the expansion valve. The heat exchanger should be located near the condenser or receiver to achieve subcooling before pressure drop occurs. • Evaporating small amounts of expected liquid refrigerant returning from evaporators in certain applications. Many heat pumps incorporating reversals of the refrigerant cycle include a suctionline accumulator and liquid-suction heat exchanger arrangement to trap liquid floodbacks and vaporize them slowly between cycle reversals. If an evaporator design makes a deliberate slight overfeed of refrigerant necessary, either to improve evaporator performance or to return oil out of the evaporator, a liquid-suction heat exchanger is needed to evaporate the refrigerant. A flooded water cooler usually incorporates an oil-rich liquid bleed from the shell into the suction line for returning oil. The liquid-suction heat exchanger boils liquid refrigerant out of the mixture in the suction line. Exchangers used for this purpose should be placed in a horizontal run near the evaporator. Several types of liquid-suction heat exchangers are used. Liquid and Suction Line Soldered Together. The simplest form of heat exchanger is obtained by strapping or soldering the suction and liquid lines together to obtain counterflow and then insulating the lines as a unit. To maximize capacity, the liquid line should always be on the bottom of the suction line, because liquid in a suction line runs along the bottom (Figure 27). This arrangement is limited by the amount of suction line available. Shell-and-Coil or Shell-and-Tube Heat Exchangers (Figure 28). These units are usually installed so that the suction outlet drains the shell. When the units are used to evaporate liquid refrigerant

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2014 ASHRAE Handbook—Refrigeration (SI)

returning in the suction line, the free-draining arrangement is not recommended. Liquid refrigerant can run along the bottom of the heat exchanger shell, having little contact with the warm liquid coil, and drain into the compressor. By installing the heat exchanger at a slight angle to the horizontal (Figure 29) with gas entering at the bottom and leaving at the top, any liquid returning in the line is trapped in the shell and held in contact with the warm liquid coil, where most of it is vaporized. An oil return line, with a metering valve and solenoid valve (open only when the compressor is running), is required to return oil that collects in the trapped shell. Concentric Tube-in-Tube Heat Exchangers. The tube-intube heat exchanger is not as efficient as the shell-and-finned-coil type. It is, however, quite suitable for cleaning up small amounts of excessive liquid refrigerant returning in the suction line. Figure 30 shows typical construction with available pipe and fittings. Plate Heat Exchangers. Plate heat exchangers provide highefficiency heat transfer. They are very compact, have low pressure drop, and are lightweight devices. They are good for use as liquid subcoolers. For air-conditioning applications, heat exchangers are recommended for liquid subcooling or for clearing up excess liquid in the suction line. For refrigeration applications, heat exchangers are recommended to increase cycle efficiency, as well as for liquid subcooling and removing small amounts of excess liquid in the suction line. Excessive superheating of the suction gas should be avoided.

Two-Stage Subcoolers To take full advantage of the two-stage system, the refrigerant liquid should be cooled to near the interstage temperature to reduce the amount of flash gas handled by the low-stage compressor. The net result is a reduction in total system power requirements. The

amount of gain from cooling to near interstage conditions varies among refrigerants. Figure 31 illustrates an open or flash-type cooler. This is the simplest and least costly type, which has the advantage of cooling liquid to the saturation temperature of the interstage pressure. One disadvantage is that the pressure of cooled liquid is reduced to interstage pressure, leaving less pressure available for liquid transport. Although the liquid temperature is reduced, the pressure drops correspondingly, and the expansion device controlling flow to the cooler must be large enough to pass all the liquid refrigerant flow. Failure of this valve could allow a large flow of liquid to the upper-stage compressor suction, which could seriously damage the compressor. Liquid from a flash cooler is saturated, and liquid from a cascade condenser usually has little subcooling. In both cases, the liquid temperature is usually lower than the temperature of the surroundings. Thus, it is important to avoid heat input and pressure losses that would cause flash gas to form in the liquid line to the expansion device or to recirculating pumps. Cold liquid lines should be insulated, because expansion devices are usually designed to feed liquid, not vapor. Figure 32 shows the closed or heat exchanger type of subcooler. It should have sufficient heat transfer surface to transfer heat from the liquid to the evaporating refrigerant with a small final temperature difference. Pressure drop should be small, so that full pressure is available for feeding liquid to the expansion device at the lowtemperature evaporator. The subcooler liquid control valve should be sized to supply only the quantity of refrigerant required for the subcooling. This prevents a tremendous quantity of liquid from flowing to the upper-stage suction in the event of a valve failure.

Discharge Line Oil Separators Oil is always in circulation in systems using halocarbon refrigerants. Refrigerant piping is designed to ensure that this oil passes

Fig. 27 Soldered Tube Heat Exchanger

Fig. 28 Shell-and-Finned-Coil Heat Exchanger Fig. 30 Tube-in-Tube Heat Exchanger

Fig. 29 Shell-and-Finned-Coil Exchanger Installed to Prevent Liquid Floodback

Fig. 31 Flash-Type Cooler

Halocarbon Refrigeration Systems through the entire system and returns to the compressor as fast as it leaves. Although well-designed piping systems can handle the oil in most cases, a discharge-line oil separator can have certain advantages in some applications (see Chapter 11), such as • In systems where it is impossible to prevent substantial absorption of refrigerant in the crankcase oil during shutdown periods. When the compressor starts up with a violent foaming action, oil is thrown out at an accelerated rate, and the separator immediately returns a large portion of this oil to the crankcase. Normally, the system should be designed with pumpdown control or crankcase heaters to minimize liquid absorption in the crankcase. • In systems using flooded evaporators, where refrigerant bleedoff is necessary to remove oil from the evaporator. Oil separators reduce the amount of bleedoff from the flooded cooler needed for operation. • In direct-expansion systems using coils or tube bundles that require bottom feed for good liquid distribution and where refrigerant carryover from the top of the evaporator is essential for proper oil removal. • In low-temperature systems, where it is advantageous to have as little oil as possible going through the low side. • In screw-type compressor systems, where an oil separator is necessary for proper operation. The oil separator is usually supplied with the compressor unit assembly directly from the compressor manufacturer. • In multiple compressors operating in parallel. The oil separator can be an integral part of the total system oil management system. In applying oil separators in refrigeration systems, the following potential hazards must be considered: • Oil separators are not 100% efficient, and they do not eliminate the need to design the complete system for oil return to the compressor. • Oil separators tend to condense out liquid refrigerant during compressor off cycles and on compressor start-up. This is true if the condenser is in a warm location, such as on a roof. During the off cycle, the oil separator cools down and acts as a condenser for refrigerant that evaporates in warmer parts of the system. A cool oil separator may condense discharge gas and, on compressor start-up, automatically drain it into the compressor crankcase. To minimize this possibility, the drain connection from the oil separator can be connected into the suction line. This line should be equipped with a shutoff valve, a fine filter, hand throttling and solenoid valves, and a sight glass. The throttling valve should be adjusted so that flow through this line is only a little greater than would normally be expected to return oil through the suction line. • The float valve is a mechanical device that may stick open or closed. If it sticks open, hot gas will continuously bypass to the compressor crankcase. If the valve sticks closed, no oil is returned to the compressor. To minimize this problem, the separator can be

Fig. 32 Closed-Type Subcooler

1.31 supplied without an internal float valve. A separate external float trap can then be located in the oil drain line from the separator preceded by a filter. Shutoff valves should isolate the filter and trap. The filter and traps are also easy to service without stopping the system. The discharge line pipe size into and out of the oil separator should be the full size determined for the discharge line. For separators that have internal oil float mechanisms, allow enough room to remove the oil float assembly for servicing. Depending on system design, the oil return line from the separator may feed to one of the following locations: • Directly to the compressor crankcase • Directly into the suction line ahead of the compressor • Into an oil reservoir or device used to collect oil, used for a specifically designed oil management system When a solenoid valve is used in the oil return line, the valve should be wired so that it is open when the compressor is running. To minimize entrance of condensed refrigerant from the low side, a thermostat may be installed and wired to control the solenoid in the oil return line from the separator. The thermostat sensing element should be located on the oil separator shell below the oil level and set high enough so that the solenoid valve will not open until the separator temperature is higher than the condensing temperature. A superheat-controlled expansion valve can perform the same function. If a discharge line check valve is used, it should be downstream of the oil separator.

Surge Drums or Accumulators A surge drum is required on the suction side of almost all flooded evaporators to prevent liquid slopover to the compressor. Exceptions include shell-and-tube coolers and similar shell-type evaporators, which provide ample surge space above the liquid level or contain eliminators to separate gas and liquid. A horizontal surge drum is sometimes used where headroom is limited. The drum can be designed with baffles or eliminators to separate liquid from the suction gas. More often, sufficient separation space is allowed above the liquid level for this purpose. Usually, the design is vertical, with a separation height above the liquid level of 600 to 750 mm and with the shell diameter sized to keep suction gas velocity low enough to allow liquid droplets to separate. Because these vessels are also oil traps, it is necessary to provide oil bleed. Although separators may be fabricated with length-to-diameter (L/D) ratios of 1/1 up to 10/1, the lowest-cost separators are usually for L/D ratios between 3/1 and 5/1.

Compressor Floodback Protection Certain systems periodically flood the compressor with excessive amounts of liquid refrigerant. When periodic floodback through the suction line cannot be controlled, the compressor must be protected against it. The most satisfactory method appears to be a trap arrangement that catches liquid floodback and (1) meters it slowly into the suction line, where the floodback is cleared up with a liquid-suction heat interchanger; (2) evaporates the liquid 100% in the trap itself by using a liquid coil or electric heater, and then automatically returns oil to the suction line; or (3) returns it to the receiver or to one of the evaporators. Figure 29 illustrates an arrangement that handles moderate liquid floodback, disposing of liquid by a combination of boiling off in the exchanger and limited bleedoff into the suction line. This device, however, does not have sufficient trapping volume for most heat pump applications or hot-gas defrost systems using reversal of the refrigerant cycle.

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 34 Drier with Piping Connections A three-valve bypass is usually used, as shown in Figure 34, to provide a way to isolate the drier for servicing. The refrigerant charging connection should be located between the receiver outlet valve and liquid-line drier so that all refrigerant added to the system passes through the drier. Reliable moisture indicators can be installed in refrigerant liquid lines to provide a positive indication of when the drier cartridge should be replaced.

Strainers

Fig. 33 Compressor Floodback Protection Using Accumulator with Controlled Bleed For heavier floodback, a larger volume is required in the trap. The arrangement shown in Figure 33 has been used successfully in reverse-cycle heat pump applications using halocarbon refrigerants. It consists of a suction-line accumulator with enough volume to hold the maximum expected floodback and a large enough diameter to separate liquid from suction gas. Trapped liquid is slowly bled off through a properly sized and controlled drain line into the suction line, where it is boiled off in a liquid-suction heat exchanger between cycle reversals. With the alternative arrangement shown, the liquid/oil mixture is heated to evaporate the refrigerant, and the remaining oil is drained into the crankcase or suction line.

Refrigerant Driers and Moisture Indicators The effect of moisture in refrigeration systems is discussed in Chapters 6 and 7. Using a permanent refrigerant drier is recommended on all systems and with all refrigerants. It is especially important on low-temperature systems to prevent ice from forming at expansion devices. A full-flow drier is always recommended in hermetic compressor systems to keep the system dry and prevent decomposition products from getting into the evaporator in the event of a motor burnout. Replaceable-element filter-driers are preferred for large systems because the drying element can be replaced without breaking any refrigerant connections. The drier is usually located in the liquid line near the liquid receiver. It may be mounted horizontally or vertically with the flange at the bottom, but it should never be mounted vertically with the flange on top because any loose material would then fall into the line when the drying element was removed.

Strainers should be used in both liquid and suction lines to protect automatic valves and the compressor from foreign material, such as pipe welding scale, rust, and metal chips. The strainer should be mounted in a horizontal line, oriented so that the screen can be replaced without loose particles falling into the system. A liquid-line strainer should be installed before each automatic valve to prevent particles from lodging on the valve seats. Where multiple expansion valves with internal strainers are used at one location, a single main liquid-line strainer will protect all of these. The liquid-line strainer can be located anywhere in the line between the condenser (or receiver) and the automatic valves, preferably near the valves for maximum protection. Strainers should trap the particle size that could affect valve operation. With pilot-operated valves, a very fine strainer should be installed in the pilot line ahead of the valve. Filter-driers dry the refrigerant and filter out particles far smaller than those trapped by mesh strainers. No other strainer is needed in the liquid line if a good filter-drier is used. Refrigeration compressors are usually equipped with a built-in suction strainer, which is adequate for the usual system with copper piping. The suction line should be piped at the compressor so that the built-in strainer is accessible for servicing. Both liquid- and suction-line strainers should be adequately sized to ensure sufficient foreign material storage capacity without excessive pressure drop. In steel piping systems, an external suction-line strainer is recommended in addition to the compressor strainer.

Liquid Indicators Every refrigeration system should have a way to check for sufficient refrigerant charge. Common devices used are liquid-line sight glass, mechanical or electronic indicators, and an external gage glass with equalizing connections and shutoff valves. A properly installed sight glass shows bubbling when the charge is insufficient. Liquid indicators should be located in the liquid line as close as possible to the receiver outlet, or to the condenser outlet if no receiver is used (Figure 35). The sight glass is best installed in a vertical section of line, far enough downstream from any valve that the resulting disturbance does not appear in the glass. If the sight glass is installed too far away from the receiver, the line pressure drop may be sufficient to cause flashing and bubbles in the

Halocarbon Refrigeration Systems

Fig. 35

Sight Glass and Charging Valve Locations

glass, even if the charge is sufficient for a liquid seal at the receiver outlet. When sight glasses are installed near the evaporator, often no amount of system overcharging will give a solid liquid condition at the sight glass because of pressure drop in the liquid line or lift. Subcooling is required here. An additional sight glass near the evaporator may be needed to check the refrigerant condition at that point. Sight glasses should be installed full size in the main liquid line. In very large liquid lines, this may not be possible; the glass can then be installed in a bypass or saddle mount that is arranged so that any gas in the liquid line will tend to move to it. A sight glass with double ports (for back lighting) and seal caps, which provide added protection against leakage, is preferred. Moisture-liquid indicators large enough to be installed directly in the liquid line serve the dual purpose of liquid-line sight glass and moisture indicator.

Oil Receivers Oil receivers serve as reservoirs for replenishing crankcase oil pumped by the compressors and provide the means to remove refrigerant dissolved in the oil. They are selected for systems having any of the following components: • Flooded or semiflooded evaporators with large refrigerant charges • Two or more compressors operated in parallel • Long suction and discharge lines • Double suction line risers A typical hookup is shown in Figure 33. Outlets are arranged to prevent oil from draining below the heater level to avoid heater burnout and to prevent scale and dirt from being returned to the compressor.

Purge Units Noncondensable gas separation using a purge unit is useful on most large refrigeration systems where suction pressure may fall below atmospheric pressure (see Figure 30 of Chapter 2).

PRESSURE CONTROL FOR REFRIGERANT CONDENSERS For more information on pressure control, see Chapter 39 of the 2012 ASHRAE Handbook—HVAC Systems and Equipment.

1.33

Fig. 36 Pressure Control for Condensers Used with Cooling Towers (Water Bypass Modulation)

Condenser-Water-Regulating Valves The shutoff pressure of the valve must be set slightly higher than the saturation pressure of the refrigerant at the highest ambient temperature expected when the system is not in operation. This ensures that the valve will not pass water during off cycles. These valves are usually sized to pass the design quantity of water at about a 170 to 200 kPa difference between design condensing pressure and valve shutoff pressure. Chapter 11 has further information.

Water Bypass In cooling tower applications, a simple bypass with a manual or automatic valve responsive to pressure change can also be used to maintain condensing pressure. Figure 36 shows an automatic threeway valve arrangement. The valve divides water flow between the condenser and the bypass line to maintain the desired condensing pressure. This maintains a balanced flow of water on the tower and pump.

Evaporative Condensers Among the methods used for condensing pressure control with evaporative condensers are (1) cycling the spray pump motor; (2) cycling both fan and spray pump motors; (3) throttling the spray water; (4) bypassing air around duct and dampers; (5) throttling air via dampers, on either inlet or discharge; and (6) combinations of these methods. For further information, see Chapter 39 of the 2012 ASHRAE Handbook—HVAC Systems and Equipment. In water pump cycling, a pressure control at the gas inlet starts and stops the pump in response to pressure changes. The pump sprays water over the condenser coils. As pressure drops, the pump stops and the unit becomes an air-cooled condenser. Constant pressure is difficult to maintain with coils of prime surface tubing because as soon as the pump stops, the pressure goes up and the pump starts again. This occurs because these coils have insufficient capacity when operating as an air-cooled condenser. The problem is not as acute with extended-surface coils. Shortcycling results in excessive deposits of mineral and scale on the tubes, decreasing the life of the water pump. One method of controlling pressure is using cycle fans and pumps. This minimizes water-side scaling. In colder climates, an indoor water sump with a remote spray pump(s) is required. The fan cycling sequence is as follows: Upon dropping pressure

Water-Cooled Condensers

• Stop fans. • If pressure continues to fall, stop pumps.

With water-cooled condensers, pressure controls are used both to maintain condensing pressure and to conserve water. On cooling tower applications, they are used only where it is necessary to maintain condensing temperatures.

• Start fans. • If pressure continues to rise, start pumps.

Upon rising pressure

1.34

2014 ASHRAE Handbook—Refrigeration (SI) The third method holds condensing pressure up by backing liquid refrigerant up in the coil to cut down on effective condensing surface. When pressure drops below the setting of the modulating control valve, it opens, allowing discharge gas to enter the liquid drain line. This restricts liquid refrigerant drainage and causes the condenser to flood enough to maintain the condenser and receiver pressure at the control valve setting. A pressure difference must be available across the valve to open it. Although the condenser imposes sufficient pressure drop at full load, pressure drop may practically disappear at partial loading. Therefore, a positive restriction must be placed parallel with the condenser and the control valve. Systems using this type of control require extra refrigerant charge. In multiple-fan air-cooled condensers, it is common to cycle fans off down to one fan and then to apply air throttling to that section or modulate the fan motor speed. Consult the manufacturer before using this method, because not all condensers are properly circuited for it. Using ambient temperature change (rather than condensing pressure) to modulate air-cooled condenser capacity prevents rapid cycling of condenser capacity. A disadvantage of this method is that the condensing pressure is not closely controlled.

Fig. 37 Pressure Control for Evaporative Condenser (Air Intake Modulation)

Microchannel Condensers The methods for low-ambient, condensing pressure control for microchannel condensers are essentially the same as those used for standard air-cooled condensers. However, because most microchannel condensers are made up of many individual heat exchangers, there is an opportunity to mechanically isolate portions of the condenser to reduce the usable surface area. This type of control scheme can be used instead of holding back excess refrigerant to flood portions of the condenser.

KEEPING LIQUID FROM CRANKCASE DURING OFF CYCLES Control of reciprocating compressors should prevent excessive accumulation of liquid refrigerant in the crankcase during off cycles. Any one of the following control methods accomplishes this.

Automatic Pumpdown Control (Direct-Expansion Air-Cooling Systems) The most effective way to keep liquid out of the crankcase during system shutdown is to operate the compressor on automatic pumpdown control. The recommended arrangement involves the following devices and provisions: Fig. 38 Pressure Control for Evaporative Condenser (Air Bypass Modulation) Damper control (Figure 37) may be incorporated in systems requiring more constant pressures (e.g., some systems using thermostatic expansion valves). One drawback of dampers is formation of ice on dampers and linkages. Figure 38 incorporates an air bypass arrangement for controlling pressure. A modulating motor, acting in response to a modulating pressure control, positions dampers so that the mixture of recirculated and cold inlet air maintains the desired pressure. In extremely cold weather, most of the air is recirculated.

Air-Cooled Condensers Methods for condensing pressure control with air-cooled condensers include (1) cycling fan motor, (2) air throttling or bypassing, (3) coil flooding, and (4) fan motor speed control. The first two methods are described in the section on Evaporative Condensers.

• A liquid-line solenoid valve in the main liquid line or in the branch to each evaporator. • Compressor operation through a low-pressure cutout providing for pumpdown whenever this device closes, regardless of whether the balance of the system is operating. • Electrical interlock of the liquid solenoid valve with the evaporator fan, so refrigerant flow stops when the fan is out of operation. • Electrical interlock of refrigerant solenoid valve with safety devices (e.g., high-pressure cutout, oil safety switch, and motor overloads), so that the refrigerant solenoid valve closes when the compressor stops. • Low-pressure control settings such that the cut-in point corresponds to a saturated refrigerant temperature lower than any expected compressor ambient air temperature. If the cut-in setting is any higher, liquid refrigerant can accumulate and condense in the crankcase at a pressure corresponding to the ambient temperature. Then, crankcase pressure would not rise high enough to reach the cut-in point, and effective automatic pumpdown would not be obtained.

Halocarbon Refrigeration Systems Crankcase Oil Heater (Direct-Expansion Systems) A crankcase oil heater with or without single (nonrecycling) pumpout at the end of each operating cycle does not keep liquid refrigerant out of the crankcase as effectively as automatic pumpdown control, but many compressors equalize too quickly after stopping automatic pumpdown control. Crankcase oil heaters maintain the crankcase oil at a temperature higher than that of other parts of the system, minimizing absorption of the refrigerant by the oil. Operation with the single pumpout arrangement is as follows. Whenever the temperature control device opens the circuit, or the manual control switch is opened for shutdown purposes, the crankcase heater is energized, and the compressor keeps running until it cuts off on the low-pressure switch. Because the crankcase heater remains energized during the complete off cycle, it is important that a continuous live circuit be available to the heater during the off time. The compressor cannot start again until the temperature control device or manual control switch closes, regardless of the position of the low-pressure switch. This control method requires • A liquid-line solenoid valve in the main liquid line or in the branch to each evaporator • Use of a relay or the maintained contact of the compressor motor auxiliary switch to obtain a single pumpout operation before stopping the compressor • A relay or auxiliary starter contact to energize the crankcase heater during the compressor off cycle and deenergize it during the compressor on cycle • Electrical interlock of the refrigerant solenoid valve with the evaporator fan, so that refrigerant flow is stopped when the fan is out of operation • Electrical interlock of refrigerant solenoid valve with safety devices (e.g., high-pressure cutout, oil safety switch, and motor overloads), so that the refrigerant flow valve closes when the compressor stops

Control for Direct-Expansion Water Chillers Automatic pumpdown control is undesirable for direct-expansion water chillers because freezing is possible if excessive cycling occurs. A crankcase heater is the best solution, with a solenoid valve in the liquid line that closes when the compressor stops.

Effect of Short Operating Cycle With reciprocating compressors, oil leaves the crankcase at an accelerated rate immediately after starting. Therefore, each start should be followed by a long enough operating period to allow the oil level to recover. Controllers used for compressors should not produce short-cycling of the compressor. Refer to the compressor manufacturer’s literature for guidelines on maximum or minimum cycles for a specified period.

HOT-GAS BYPASS ARRANGEMENTS Most large reciprocating compressors are equipped with unloaders that allow the compressor to start with most of its cylinders unloaded. However, it may be necessary to further unload the compressor to (1) reduce starting torque requirements so that the compressor can be started both with low-starting-torque prime movers and on lowcurrent taps of reduced voltage starters and (2) allow capacity control down to 0% load conditions without stopping the compressor.

Full (100%) Unloading for Starting Starting the compressor without load can be done with a manual or automatic valve in a bypass line between the hot-gas and suction lines at the compressor.

1.35 To prevent overheating, this valve is open only during the starting period and closed after the compressor is up to full speed and full voltage is applied to the motor terminals. In the control sequence, the unloading bypass valve is energized on demand of the control calling for compressor operation, equalizing pressures across the compressor. After an adequate delay, a timing relay closes a pair of normally open contacts to start the compressor. After a further time delay, a pair of normally closed timing relay contacts opens, deenergizing the bypass valve.

Full (100%) Unloading for Capacity Control Where full unloading is required for capacity control, hot-gas bypass arrangements can be used in ways that will not overheat the compressor. In using these arrangements, hot gas should not be bypassed until after the last unloading step. Hot-gas bypass should (1) give acceptable regulation throughout the range of loads, (2) not cause excessive superheating of the suction gas, (3) not cause any refrigerant overfeed to the compressor, and (4) maintain an oil return to the compressor. Hot-gas bypass for capacity control is an artificial loading device that maintains a minimum evaporating pressure during continuous compressor operation, regardless of evaporator load. This is usually done by an automatic or manual pressure-reducing valve that establishes a constant pressure on the downstream side. Four common methods of using hot-gas bypass are shown in Figure 39. Figure 39A illustrates the simplest type; it will dangerously overheat the compressor if used for protracted periods of time. Figure 39B shows the use of hot-gas bypass to the exit of the evaporator. The expansion valve bulb should be placed at least 1.5 m downstream from the bypass point of entrance, and preferably further, to ensure good mixing. In Figure 39D, the hot-gas bypass enters after the evaporator thermostatic expansion valve bulb. Another thermostatic expansion valve supplies liquid directly to the bypass line for desuperheating. It is always important to install the hot-gas bypass far enough back in the system to maintain sufficient gas velocities in suction risers and other components to ensure oil return at any evaporator loading. Figure 39C shows the most satisfactory hot-gas bypass arrangement. Here, the bypass is connected into the low side between the expansion valve and entrance to the evaporator. If a distributor is used, gas enters between the expansion valve and distributor. Refrigerant distributors are commercially available with side inlet connections that can be used for hot-gas bypass duty to a certain extent. Pressure drop through the distributor tubes must be evaluated to determine how much gas can be bypassed. This arrangement provides good oil return. Solenoid valves should be placed before the constant-pressure bypass valve and before the thermal expansion valve used for liquid injection desuperheating, so that these devices cannot function until they are required. Control valves for hot gas should be close to the main discharge line because the line preceding the valve usually fills with liquid when closed. The hot-gas bypass line should be sized so that its pressure loss is only a small percentage of the pressure drop across the valve. Usually, it is the same size as the valve connections. When sizing the valve, consult a control valve manufacturer to determine the minimum compressor capacity that must be offset, refrigerant used, condensing pressure, and suction pressure. When unloading (Figure 39C), pressure control requirements increase considerably because the only heat delivered to the condenser is that caused by the motor power delivered to the compressor. Discharge pressure should be kept high enough that the hot-gas bypass valve can deliver gas at the required rate. The condenser pressure control must be capable of meeting this condition.

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 39

Hot-Gas Bypass Arrangements

MINIMIZING REFRIGERANT CHARGE IN COMMERCIAL SYSTEMS Preventing refrigerant leaks is the most effective way to reduce halocarbons’ environmental effects. However, if a leak does occur, the consequence are reduced if the system charge has been minimized. There are many ways to reduce charge, but most require significant system modifications; consequently, charge reduction is usually performed during system remodeling or replacement. One of the best opportunities to reduce refrigerant charge exists in the distribution piping that feeds liquid to the evaporator from the receiver and returns the suction gas to the compressor. Systems serving numerous evaporators across a facility (e.g., in supermarkets) rely on a network of distribution piping, which can contain a large portion of the entire system charge. For systems that use single circuiting, in which each evaporator (or small group of adjacent evaporators) has its own liquid and suction line piped back to the compressor, charge can be significantly reduced by zoning the loads. For loads operating at similar evaporator pressures, one suction and liquid line can run from the machinery room and branch out closer to the load to feed multiple evaporators (loop piping). Expansion, solenoid, and evaporator pressure regulating valves must be next to the heat load in these systems, but benefits beyond reduced charge include cheaper installation cost and less physical space required to run the lines. Note that using hot-gas defrost with this type of piping scheme is typically not preferred, because it requires a third branched line that must also be field installed. In the liquid feed lines, subcooling the liquid can further reduce charge. Because subcooling increases the refrigerant’s quality, the

required mass flow rate is reduced, thus allowing use of smaller liquid lines (and thus smaller-volume refrigerant charges) with comparable velocities and pressure drops. Subcooling is typically chosen for its energy benefits and is also often used to protect liquid from flashing before it reaches the expansion valve, so the fact that the refrigerant charge can be reduced is often considered a secondary benefit. The other factor affecting the amount of refrigerant in the distribution piping is the equipment location. Minimizing the distance between the receiver and the evaporators also reduces the refrigerant charge in the liquid piping. For this reason, some users install compressor systems throughout their facilities instead of centralizing them in a compressor room. Distributed systems typically use quieter scroll compressors, along with special noise-reducing enclosures to allow installations in more exposed and occupied areas. Replacing or retrofitting a direct system to an indirect (or secondary) system is another way to reduce refrigerant charge in distribution piping. This method requires a much more dramatic change to the system, but it is probably the most effective because it can restrict the halocarbon refrigerant to a compact unit composed of a compressor, condenser, and evaporator. The secondary fluid can then be pumped through air-cooling heat exchangers at the load. In this type of system, only a few evaporators are required and the distribution piping is eliminated, so the chance of refrigerant leaks is dramatically reduced. Opportunities to reduce charge also exist on the high-pressure side of the system between the compressor and the receiver. In comparison to standard air-cooled condensers, systems that use watercooled condensers operate with a lower charge. If a condensing

Halocarbon Refrigeration Systems water source is available, a flat-plate condenser can be mounted near the compressors and used to reject heat from the high-pressure side of the system to the water loop. Typically, bypass lines, variablespeed pumps, and/or flow-restricting valves are used to maintain minimum condensing pressures in water-cooled condensers. Because condenser flooding is no longer required, refrigerant charge can be reduced. Microchannel condensers also have lower charges than standard air-cooled condensers but may require long runs of liquid piping in installations with indoor compressors. In systems that require flooding, microchannel condensers allow for reduced refrigerant charge because of their smaller internal volume. Alternatively, in low ambient conditions, in conjunction with fan controls, entire banks of some microchannel condensers can be isolated using solenoid valves if the outlet piping is correctly trapped; this approach provides the same benefit as condenser flooding, but requires less refrigerant.

REFRIGERANT RETROFITTING Because of the halocarbon phaseout, many users are retrofitting existing systems to use newer, more acceptable refrigerants (e.g., converting from the HCFC R-22 to the HFC R-407A). Such conversions require planning and preparation. The most glaring concern is the effect of the new refrigerant on system capacity. Not only should the capacity of the compressor(s) be considered, but also the capacity of every other component in the system (condensers, evaporators, valves, etc.). Equipment and component manufacturers often can provide the needed derating factors to adjust capacities appropriately. Before any work begins, it is a good idea to record how the system is performing: data such as highand low-side pressures and temperatures help to suggest how the system should operate after the retrofit. Any available energy data should also be recorded, so the system’s efficiency can be compared to expectations. Because pressure-temperature relationships will be different for the new refrigerant, the contractor must be prepared to adjust all the pressure controls and/or modify controller set points throughout the system. Thermal expansion valves (TXV) require attention in any retrofit. At the very least, the superheats need to be adjusted; often, the temperature-sensing bulbs and nozzles must be changed out. The designer should consult with the valve manufacturer to decide what action should be taken, and whether the entire TXV should be replaced. Changing out the system’s lubricating oil is also often required during a retrofit. Mineral oils and alkylbenzene oils are often replaced with POE oils to maintain oil miscibility with the new refrigerant. It is always important to follow a thorough change-out procedure to ensure that all traces of the existing oil are removed from the system. A typical procedure includes (among other tasks) draining the existing oil; changing out liquid driers, suction filters, and oil filters; and recharging the system with the new oil. The draining and recharging steps may need to be repeated more than once to achieve the desired purity for the new oil. Traditionally, 95% or higher purity is required. Elastomeric gasket and seal materials in the system will also react differently to new refrigerants and oils. Swell characteristics of different elastomers can be referenced from Table 9 in Chapter 29 of the 2013 ASHRAE Handbook—Fundamentals; however, testing is necessary to know exactly how gaskets and seals will react to mixtures of different refrigerants and oils and what factors other than swell may come into play, such as the overall integrity and functionality of the material. For this reason, it is common practice to change out all elastomeric gaskets and seals as part of the retrofit procedure. After the system is up and running with the new refrigerant and oil, the performance of the system can be evaluated to determine

1.37 whether it is performing as expected. Refrigerant and oil levels should also be monitored until the correct levels are achieved, and filters should be changed until the system is clean. Finally, it is always crucial to make the appropriate signage and labeling modifications to prevent anyone from topping off the system with the old refrigerant or oil.

TEMPERATURE GLIDE It is not uncommon to retrofit existing systems from single-component (azeotropic) refrigerants to blended, zeotropic refrigerants. (Refer to Chapter 2 of the 2013 ASHRAE Handbook—Fundamentals for more information on zeotropic refrigerants.) In these scenarios, the designer and contractor needs to be aware of how the refrigerant’s temperature glide behaves throughout the system and know how to properly use the bubble, mean, and dew-point temperatures at the evaporator and condenser to accurately calculate subcooling and superheating. Beyond this, the designer must know what temperatures to use to properly size equipment. When a zeotropic refrigerant starts to condense in the condenser, it does so at a constant pressure (ignoring pressure drop) at the refrigerant’s dew-point temperature. As the refrigerant continues to condense, its temperature drops until it reaches the bubble-point temperature, at which point it is fully condensed. The liquid can then be subcooled. Conversely, when the refrigerant starts to boil in the evaporator, it starts at the bubble-point temperature and is not fully evaporated until it reaches the dew-point temperature. The gas can then be superheated. So, when calculating subcooling at the condenser exit, the bubble-point temperature must represent the saturation point; when calculating superheating at the evaporator exit, the dew-point temperature must represent the saturation point. Refrigerant manufacturers publish pressure-temperature charts that allow the bubble, mean, and dew-point temperatures to be easily referenced given a specific pressure. This temperature change behavior during the phase-change process is known as the refrigerant’s temperature glide and is caused by the varying boiling points of the constituent refrigerants within the mixture. Blended refrigerants essentially separate (fractionate) during phase changes, so leaky condensers and evaporators create concern: refrigerant composition changes can occur in the system, leading to unpredictable system operation. For this reason, it is necessary to only charge systems with refrigerant in the liquid state unless the entire cylinder will be immediately used. Furthermore, if a leak occurs and the system is repaired, the refrigerant composition should be checked for significant changes before topping off the system. When calculating temperature differences to check the rated capacity of existing condensers and evaporators, the mean temperatures should be used along with any derating factors provided by the manufacturer. When checking the capacity of existing compressors, however, using mean temperatures yields a slightly smaller capacity than they actually have because ANSI/AHRI Standard 540-2004 requires, when rating compressor capacities, that dewpoint temperatures be used as the reference temperatures at the corresponding evaporating and condensing pressures. The challenge, however, exists in accurately determining the dew-point temperatures. Simply adding half of the glide to the mean temperature may not be accurate: it is difficult to determine what the actual mean temperature really must be for effective evaporator or condenser operation. Because most published capacity data for heat exchangers are based on temperature that is assumed to be constant during phase change, supplemental derating factors must be used.

1.38

2014 ASHRAE Handbook—Refrigeration (SI) REFERENCES

AHRI. 2004. Performance rating of positive displacement refrigerant compressors and compressor units. ANSI/AHRI Standard 540-2004. AirConditioning, Heating, and Refrigeration Institute, Arlington, VA. Alofs, D.J., M.M. Hasan, and H.J. Sauer, Jr. 1990. Influence of oil on pressure drop in refrigerant compressor suction lines. ASHRAE Transactions 96:1. ASHRAE. 2013. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2013. ASME. 2006. Refrigeration piping and heat transfer components. ANSI/ ASME Standard B31.5-2006. American Society of Mechanical Engineers, New York. Atwood, T. 1990. Pipe sizing and pressure drop calculations for HFC-134a. ASHRAE Journal 32(4):62-66. Calm, J.M. 2008. The next generation of refrigerants—Historical review, considerations, and outlook. Ecolibrium Nov.:24-33. Colebrook, D.F. 1938, 1939. Turbulent flow in pipes. Journal of the Institute of Engineers 11.

Cooper, W.D. 1971. Influence of oil-refrigerant relationships on oil return. ASHRAE Symposium Bulletin PH71(2):6-10. Giunta, C.J. 2006. Thomas Midgley, Jr. and the invention of chlorofluorocarbon refrigerants: It ain’t necessarily so. Bulletin for the History of Chemistry 31(2):66-74. IPCC. 1990. First Assessment Report (FAR) Overview Chapter. Cambridge University Press Jacobs, M.L., F.C. Scheideman, F.C. Kazem, and N.A. Macken. 1976. Oil transport by refrigerant vapor. ASHRAE Transactions 81(2):318-329. Keating, E.L., and R.A. Matula. 1969. Correlation and prediction of viscosity and thermal conductivity of vapor refrigerants. ASHRAE Transactions 75(1). Stoecker, W.F. 1984. Selecting the size of pipes carrying hot gas to defrosted evaporators. International Journal of Refrigeration 7(4):225-228. Timm, M.L. 1991. An improved method for calculating refrigerant line pressure drops. ASHRAE Transactions 97(1):194-203. Wile, D.D. 1977. Refrigerant line sizing. ASHRAE.

CHAPTER 2

AMMONIA REFRIGERATION SYSTEMS EQUIPMENT............................................................................. 2.1 Compressors............................................................................... 2.1 Condensers................................................................................. 2.5 Evaporators................................................................................ 2.8 Evaporator Piping ..................................................................... 2.9 Vessels ...................................................................................... 2.11 Piping....................................................................................... 2.14

Controls .................................................................................... SYSTEMS.................................................................................. Single-Stage Systems ................................................................ Economized Systems................................................................. Multistage Systems ................................................................... Liquid Recirculation Systems ................................................... Safety Considerations...............................................................

C

chlorofluorocarbon (CFC) and hydrochlorofluorocarbon (HCFC) refrigerants. Ammonia secondary systems that circulate chilled water or another secondary refrigerant are a viable alternative to halocarbon systems, although ammonia is inappropriate for direct refrigeration systems (ammonia in the air unit coils) for HVAC applications. Ammonia packaged chilling units are available for HVAC applications. As with the installation of any air-conditioning unit, all applicable codes, standards, and insurance requirements must be followed.

USTOM-ENGINEERED ammonia (R-717) refrigeration systems often have design conditions that span a wide range of evaporating and condensing temperatures. Examples are (1) a food freezing plant operating from 10 to –45°C; (2) a candy storage requiring 15°C db with precise humidity control; (3) a beef chill room at –2 to –1°C with high humidity; (4) a distribution warehouse requiring multiple temperatures for storing ice cream, frozen food, meat, and produce and for docks; and (5) a chemical process requiring multiple temperatures ranging from 15 to –50°C. Ammonia is the refrigerant of choice for many industrial refrigeration systems. See Chapter 24 for information on refrigeration load calculations. The figures in this chapter are for illustrative purposes only, and may not show all the required elements (e.g., valves). For safety and minimum design criteria for ammonia systems, refer to ASHRAE Standard 15, IIAR Bulletin 109, IIAR Standard 2, and applicable state and local codes.

History of Ammonia Refrigeration First synthesized in 1823, ammonia was first used as a refrigerant in an ice-making vapor absorption system developed by Ferdinand Carré, a French engineer and inventor, in 1858 (GPO 1893). The Carré machine used an aqueous ammonia solution, with water as the absorbent and ammonia as the refrigerant. This type of vapor absorption system remains in use today. Use of ammonia as a refrigerant in vapor compression systems followed. David Boyle established an ice production plant in Jefferson, TX, in 1873 using an improved compressor design, and he later set up the Boyle Ice Machine Co. in Chicago, IL, in 1878 (Balmer 2010; Woolrich et al. n.d.). With the financial backing of several breweries, Professor Carl von Linde of Munich, Germany, had 30 ice machines of his design built between 1875 and 1881 (Dincer 1997; Schmidt 1908). The first commercial production of synthetic ammonia began in 1913 (IIAR n.d.). Worldwide annual production of ammonia is approximately 135 million metric tons, of which 9.4 million metric tons was produced in the United States in 2011 (USGS 2012). Over 80% of the ammonia produced is used in agriculture as fertilizer; less than 2% is used as a refrigerant (ASHRAE 2002). Of the three primary first-generation refrigerants used during the 1920s [i.e., ammonia (R-717), chloromethane (R-40), and sulfur dioxide (R-764)], only ammonia remains in use today as a refrigerant. Ammonia is considered a natural refrigerant because it is a common, naturally occurring compound, and it naturally breaks down into hydrogen and nitrogen.

Ammonia Refrigerant for HVAC Systems There is renewed interest in using ammonia for HVAC systems, in part because of the scheduled phaseout and increasing costs of The preparation of this chapter is assigned to TC 10.3, Refrigerant Piping, Controls and Accessories.

2.1

2.15 2.18 2.18 2.18 2.19 2.20 2.25

EQUIPMENT COMPRESSORS Compressors available for single- and multistage applications include the following: • Rotary vane • Reciprocating • Rotary screw Rotary vane compressors are typically used for low-stage (booster) compressor applications. Reciprocating and screw compressors can be used as single-stage, low-stage (booster), or high-stage machines and can also be internally compounded to provide multiple compression stages on one compressor body. The reciprocating compressor is the most common compressor used in small, 75 kW or less, single-stage or multistage systems. The screw compressor is the predominant compressor above 75 kW, in both single- and multistage systems. Various combinations of compressors may be used in multistage systems. Rotary vane and screw compressors are frequently used for the low-pressure stage, where large volumes of gas must be moved. The high-pressure stage may be a reciprocating or screw compressor. When selecting a compressor, consider the following: • System size and capacity requirements. • Location, such as indoor or outdoor installation at ground level or on the roof. • Equipment noise. • Part- or full-load operation. • Winter and summer operation. • Pulldown time required to reduce the temperature to desired conditions for either initial or normal operation. The temperature must be pulled down frequently for some applications for a process load, whereas a large cold-storage warehouse may require pulldown only once in its lifetime. Lubricant Cooling. When a reciprocating compressor requires lubricant cooling, an external heat exchanger using a refrigerant or secondary cooling is usually added. Screw compressor lubricant cooling is covered in detail in the section on Screw Compressors. Compressor Drives. The correct electric motor size(s) for a multistage system is determined by pulldown load. When the final

2.2

2014 ASHRAE Handbook—Refrigeration (SI)

low-stage operating level is –75°C, the pulldown load can be three times the operating load. Positive-displacement reciprocating compressor motors are usually selected for about 150% of operating power requirements for 100% load. The compressor’s unloading mechanism can be used to prevent motor overload. Electric motors should not be overloaded, even when a service factor is indicated. For screw compressor applications, motors should be sized by adding 10% to the operating power. Screw compressors have built-in unloading mechanisms to prevent motor overload. The motor should not be oversized, because an oversized motor has a lower power factor and lower efficiency at design and reduced loads. Steam turbines or gasoline, natural gas, propane, or diesel internal combustion engines are used when electricity is unavailable, or if the selected energy source is cheaper. Sometimes they are used in combination with electricity to reduce peak demands. The power output of a given engine size can vary as much as 15% depending on the fuel selected. Steam turbine drives for refrigerant compressors are usually limited to very large installations where steam is already available at moderate to high pressure. In all cases, torsional analysis is required to determine what coupling must be used to dampen out any pulsations transmitted from the compressor. For optimum efficiency, a turbine should operate at a high speed that must be geared down for reciprocating and possibly screw compressors. Neither the gear reducer nor the turbine can tolerate a pulsating backlash from the driven end, so torsional analysis and special couplings are essential. Advantages of turbines include variable speed for capacity control and low operating and maintenance costs. Disadvantages include higher initial costs and possible high noise levels. The turbine must be started manually to bring the turbine housing up to temperature slowly and to prevent excess condensate from entering the turbine. The standard power rating of an engine is the absolute maximum, not the recommended power available for continuous use. Also, torque characteristics of internal combustion engines and electric motors differ greatly. The proper engine selection is at 75% of its maximum power rating. For longer life, the full-load speed should be at least 10% below maximum engine speed. Internal combustion engines, in some cases, can reduce operating cost below that for electric motors. Disadvantages include (1) higher initial cost of the engine, (2) additional safety and starting controls, (3) higher noise levels, (4) larger space requirements, (5)

air pollution, (6) requirement for heat dissipation, (7) higher maintenance costs, and (8) higher levels of vibration than with electric motors. A torsional analysis must be made to determine the proper coupling if engine drives are chosen.

Reciprocating Compressors Piping. Figure 1 shows a typical piping arrangement for two compressors operating in parallel off the same suction main. Suction mains should be laid out with the objective of returning only clean, dry gas to the compressor. This usually requires a suction trap sized adequately for gravity gas and liquid separation based on permissible gas velocities for specific temperatures. A dead-end trap can usually trap only scale and lubricant. As an alternative, a shelland-coil accumulator with a warm liquid coil may be considered. Suction mains running to and from the suction trap or accumulator should be pitched toward the trap at 10 mm per metre for liquid drainage. It is also good practice to connect compressor suction branch piping above the centerline of the suction main. In sizing suction mains and takeoffs from mains to compressors, consider how pressure drop in the selected piping affects the compressor size required. First costs and operating costs for compressor and piping selections should be optimized. Good suction line systems have a total friction drop of 0.5 to 1.5 K pressure drop equivalent. Practical suction line friction losses should not exceed 0.01 K equivalent per metre equivalent length. A well-designed discharge main has a total friction loss of 7 to 14 kPa. Generally, a slightly oversized discharge line is desirable to hold down discharge pressure and, consequently, discharge temperature and energy costs. Where possible, discharge mains should be pitched (10 mm/m) toward the condenser without creating a liquid trap; otherwise, pitch should be toward the discharge line separator. High- and low-pressure cutouts and gages and lubricant pressure failure cutout are installed on the compressor side of the stop valves to protect the compressor. Lubricant Separators. Lubricant separators are located in the discharge line of each compressor (Figure 1A). A high-pressure float valve drains lubricant back into the compressor crankcase or lubricant receiver. The separator can be placed away from the compressor, so any extra pipe length can be used to cool the discharge gas before it enters the separator. This reduces the temperature of the ammonia vapor and makes the separator more effective. A discharge

Fig. 1 Schematic of Reciprocating Compressors Operating in Parallel

Ammonia Refrigeration Systems gas heat exchanger may also be used to cool the gas before it enters the oil separator. Liquid ammonia must not reach the crankcase. Often, a valve (preferably automatic) is installed in the drain from the lubricant separator, open only when the temperature at the bottom of the separator is higher than the condensing temperature. Some manufacturers install a small electric heater at the bottom of a vertical lubricant trap instead. The heater is actuated when the compressor is not operating. Separators exposed to cold must be insulated to prevent ammonia condensation. Venting the high-pressure gas in the oil separator to the crankcase or suction line also helps prevent ammonia condensation. A filter is recommended in the drain line on the downstream side of the high-pressure float valve. Lubricant Receivers. Figure 1B illustrates two compressors on the same suction line with one discharge-line lubricant separator. The separator float drains into a lubricant receiver, which maintains a reserve supply of lubricant for the compressors. Compressors should be equipped with crankcase floats to regulate lubricant flow to the crankcase. Discharge Check Valves and Discharge Lines. Discharge check valves on the downstream side of each lubricant separator prevent high-pressure gas from flowing into an inactive compressor and causing condensation (Figure 1A). The discharge line from each compressor should enter the discharge main at a 45° maximum angle in the horizontal plane so the gas flows smoothly. Unloaded Starting. Unloaded starting is frequently needed to stay within the torque or current limitations of the motor. Most compressors are unloaded either by holding each cylinder’s suction valve open or by external bypassing. Control can be manual or automatic. Suction Gas Conditioning. Suction main piping should be insulated, complete with vapor retarder to minimize thermal losses, to prevent sweating and/or ice build-up on the piping, and to limit superheat at the compressor. Additional superheat increases discharge temperatures and reduces compressor capacity. Low discharge temperatures in ammonia plants are important to reduce lubricant carryover and because compressor lubricant can carbonize at higher temperatures, which can cause cylinder wall scoring and lubricant sludge throughout the system. Discharge temperatures above 120°C should be avoided at all times. Lubricants should have flash-point temperatures above the maximum expected compressor discharge temperature. Cooling. Generally, ammonia compressors are constructed with internally cast cooling passages along the cylinders and/or in the top heads. These passages provide space for circulating a heat transfer medium, which minimizes heat conduction from the hot discharge gas to the incoming suction gas and lubricant in the compressor’s crankcase. An external lubricant cooler is supplied on most reciprocating ammonia compressors. Water is usually the medium circulated through these passages (water jackets) and the lubricant cooler at a rate of about 2 mL/s per kilowatt of refrigeration. Lubricant in the crankcase (depending on type of construction) is about 50°C. Temperatures above this level reduce the lubricant’s lubricating properties. For compressors operating in ambients above 0°C, water flow controlled by a solenoid valve in the inlet line is desirable to automate the system and prevent any refrigerant condensing above the pistons. When the compressor stops, water flow must be stopped to keep residual gas from condensing and to conserve water. A waterregulating valve, installed in the water supply line with the sensing bulb in the water return line, is also recommended. This type of cooling is shown in Figure 2. The thermostat in the water line leaving the jacket serves as a safety cutout to stop the compressor if the temperature becomes too high.

2.3 For compressors where ambient temperatures may be below 0°C, a way to drain the jacket on shutdown to prevent freeze-up must be provided. One method is shown in Figure 3. Water flow is through the inlet line normally closed solenoid valve, which is energized when the compressor starts. Water then circulates through the lubricant cooler and the jacket, and out through the water return line. When the compressor stops, the solenoid valve in the water inlet line is deenergized and stops water flow to the compressor. At the same time, the drain line normally open solenoid valve deenergizes and opens to drain the water out of the low point to wastewater treatment. The check valves in the air vent lines open when pressure is relieved and allow the jacket and cooler to be drained. Each flapper check valve is installed so that water pressure closes it, but absence of water pressure allows it to swing open. For compressors in spaces below 0°C or where water quality is very poor, cooling is best handled by using an inhibited glycol solution or other suitable fluid in the jackets and lubricant cooler and cooling with a secondary heat exchanger. This method for cooling reciprocating ammonia compressors eliminates fouling of the lubricant cooler and jacket normally associated with city water or cooling tower water.

Rotary Vane, Low-Stage Compressors Piping. Rotary vane compressors have been used extensively as low-stage compressors in ammonia refrigeration systems. Now, however, the screw compressor has largely replaced the rotary vane compressor for ammonia low-stage compressor applications. Piping requirements for rotary vane compressors are the same as for reciprocating compressors. Most rotary vane compressors are lubricated by injectors because they have no crankcase. In some designs,

Fig. 2

Jacket Water Cooling for Ambient Temperatures Above Freezing

Fig. 3 Jacket Water Cooling for Ambient Temperatures Below Freezing

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2014 ASHRAE Handbook—Refrigeration (SI)

a lubricant separator, lubricant receiver, and cooler are required on the discharge of these compressors; a pump recirculates lubricant to the compressor for both cooling and lubrication. In other rotary vane compressor designs, a discharge lubricant separator is not used, and lubricant collects in the high-stage suction accumulator or intercooler, from which it may be drained. Lubricant for the injectors must periodically be added to a reservoir. Cooling. The compressor jacket is cooled by circulating a cooling fluid, such as a water/glycol solution or lubricant. Lubricant is recommended, because it will not freeze and can serve both purposes (Figure 4).

Screw Compressors Helical screw compressors are the choice for most industrial refrigeration systems. All helical screw compressors have a constantvolume (displacement) design. The volume index Vi refers to the internal volume ratio of the compressor (i.e., the reduction in volume of the compressed gas from suction to discharge of the compressor). Capacity control is accomplished by use of a slide valve, bypass ports, or by controlling the speed [variable-frequency drive (VFD)]. The slide valve and bypass ports control capacity by only using a portion of the screw(s) for the compression process. Some compressors are designed with a fixed Vi. When Vi is fixed, the compressor functions most efficiently at a certain compression ratio (CR). In selecting a fixed Vi compressor, consider the average CR rather than the maximum CR. A guide to proper compressor selection is based on the equation Vik = CR, where k = 1.4 for ammonia. For example, a screw compressor operating at 265 kPa suction and 1350 kPa discharge has a CR = 5.09. Therefore, Vi = 3.2 (Vi = CR1/k). Thus, a compressor with the Vi at or close to 3.2 is the best selection. Because ambient conditions vary throughout the year, the average condensing temperature may be 24°C (969 kPa). With the lower discharge pressure, the average compressor CR is 3.65 and the ideal Vi is 2.52. Therefore, a compressor with the Vi at or close to 2.5 is the proper selection to optimize efficiency. Some compressors are equipped with a variable Vi control. This makes compressor selection simpler, because the volume index can vary for different operating conditions. Therefore, the internal compression ratio can automatically match the external pressure ratio. Typically, screw compressors with variable Vi can control between 2.2 and 5.0 Vi. Variable-Vi compressors are beneficial over a wide

Fig. 4 Rotary Vane Booster Compressor Cooling with Lubricant

range of system pressure ratios to improve efficiency as condensing pressures vary. Piping. Oil-flooded screw compressors are the most common type of screw compressor used in refrigeration. Introduced in the late 1950s as an alternative to dry compressors with a symmetric rotor profile, oil-flooded compressors rapidly gained acceptance in many conventional reciprocating and small centrifugal applications. These compressors typically have oil supplied to the compression area at a volume rate of about 0.5% of the displacement volume. Some of this oil is used for lubricating the bearings and shaft seal. Typically, paraffinic or naphthenic mineral oils are used, though synthetics are being used more frequently on some applications. The oil fulfills three primary purposes: sealing, cooling, and lubrication. The oil tends to fill any leakage paths between and around the screws. This provides a good volumetric efficiency even at high compression ratios. Normal volumetric efficiency exceeds 85% even with a compression ratio of 25. The oil sealing also helps maintain good volumetric efficiency with decreased operating speeds. The cooling function of the oil transfers much of the heat of compression from the gas to the oil, keeping typical discharge temperatures below 90°C. This allows high compression ratios without the danger of oil breakdown. The oil’s lubrication function protects the bearings, seals, and screw contact areas. Oil injection to the screw compressor is normally achieved by one of two methods: • An oil pump operates and builds pressure over compressor discharge pressure for oil injection. The pump may be required when the screw compressor is operating at a low compression ratio or if the compressor bearing design requires oil pressure greater than compressor discharge pressure. • Operation without a pump relies on the differential pressure across the screw compressor as the driving force for the oil injection. Some screw compressors may use a combination of both methods to achieve proper oil injection. The pump may only operate for a period of time when the compression ratio is below a set value. This option is shown schematically in Figure 5. Oil injection requires an oil separator to remove the oil from the high-pressure refrigerant. Oil separators are designed to satisfy requirements of system type, refrigerant, and heat transfer equipment being used. Modern separation equipment routinely limits oil carryover to the refrigeration system to less than 5 mg/kg of oil in proportion to the circulated refrigerant.

Fig. 5 Screw Compressor Flow Diagram with Optional Oil Pump

Ammonia Refrigeration Systems

Fig. 6 Screw Compressor Flow Diagram with Liquid Injection Oil Cooling Because the oil absorbs a significant amount of the heat of compression in oil-flooded operation, oil cooling must be used to maintain low discharge temperatures. Common oil-cooling methods include the following: • Direct injection of liquid refrigerant into the screw compression process. The injected liquid refrigerant amount is normally controlled by sensing the compressor discharge temperature. The refrigerant is modulated with a thermal expansion valve to maintain a constant discharge temperature. Some of the injected liquid mixes with the oil and reduces the amount of internal volume available for suction gas to the compressor. Therefore, the compressor capacity is reduced. In addition, the liquid absorbs heat and expands to vapor, which requires additional power to compress. Screw compressors are normally designed with the liquid injection ports as late as possible in the compression process, to minimize the capacity and power penalties. Refrigerant liquid for liquid-injection oil cooling must come from a dedicated supply. The source may be the system receiver or a separate receiver; a 5 min uninterrupted supply of refrigerant liquid is usually adequate. Refrigerant injection cooling is shown schematically in Figure 6. Depending on the application, this cooling method usually decreases compressor efficiency and capacity but lowers equipment cost. • External water or glycol heat exchangers for oil cooling. With this configuration, heat is removed from the oil by using an external oil cooler. Cooling tower water, a separate evaporative cooler, underfloor glycol, and various other sources of water or glycol are used to circulate through the oil cooler heat exchanger and remove the heat of compression. A three-way oil temperature control valve is typically used in the compressor oil piping to control oil temperature. This method of oil cooling does not affect compressor efficiency or capacity. The external heat exchanger for oil cooling is shown schematically in Figure 7. • External refrigerant heat exchanger for oil cooling (thermosiphon). With this configuration, heat is removed from the oil by using an external oil cooler and high-pressure liquid refrigerant from the system. Indirect or thermosiphon lubricant cooling for low-stage screw compressors rejects the lubricant cooling load to the condenser or auxiliary cooling system; this load is not transferred to the high-stage compressor, thus improving system efficiency. Thermosiphon lubricant cooling is the most common method of oil cooling in refrigeration. In this system, highpressure refrigerant liquid from the condenser, which boils at

2.5

Fig. 7 Screw Compressor Flow Diagram with External Heat Exchanger for Oil Cooling condensing temperature/pressure (usually 32 to 35°C design), cools lubricant in a heat exchanger. The thermosiphon oil cooler is also shown schematically in Figure 7. A typical thermosiphon oilcooling system with multiple heat exchangers is shown schematically in Figure 8. Note that the refrigerant liquid supply to the oil cooler receives priority over the feed to the system low side. It is important that the gas equalizing line (vent) off the top of the thermosiphon receiver be adequately sized to match the oil cooler load to prevent the thermosiphon receiver from becoming gas bound. It is also good practice to slope the two-phase flow return line from the oil cooler to the thermosiphon vessel down in the direction of flow at 20 mm/m. A three-way oil control valve may also be used to control oil temperature to the compressor.

CONDENSERS Condensers should be selected on the basis of total heat rejection at maximum load. Often, the heat rejected at the start of pulldown is several times the amount rejected at normal, low-temperature operating conditions. Some means, such as compressor unloading, can be used to limit the maximum amount of heat rejected during pulldown. If the condenser is not sized for pulldown conditions, and compressor capacity cannot be limited during this period, condensing pressure might increase enough to shut down the system.

Condenser and Receiver Piping Properly designed piping around the condensers and receivers keeps the condensing surface at its highest efficiency by draining liquid ammonia out of the condenser as soon as it condenses and keeping air and other noncondensables purged. Horizontal Shell-and-Tube Condenser and Through-Type Receiver. Figure 9 shows a horizontal water-cooled condenser draining into a through (top inlet) receiver. Ammonia plants do not always require controlled water flow to maintain pressure. Usually, pressure is adequate to force the ammonia to the various evaporators without water regulation. Each situation should be evaluated by comparing water costs with input power cost savings at lower condenser pressures. Water piping should be arranged so that condenser tubes are always filled with water. Air vents should be provided on condenser heads and should have hand valves for manual purging. Receivers must be below the condenser so that the condensing surface is not flooded with ammonia. The piping should provide (1) free drainage from the condenser and (2) static height of

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 8

Fig. 9

Thermosiphon System with Receiver Mounted Above Oil Cooler

Horizontal Condenser and Top Inlet Receiver Piping

ammonia above the first valve out of the condenser greater than the pressure drop through the valve. The drain line from condenser to receiver is designed on the basis of 0.5 m/s maximum velocity to allow gas equalization between condenser and receiver. Refer to Table 2 for sizing criteria. Parallel Horizontal Shell-and-Tube Condensers. Figure 10 shows two condensers operating in parallel with one through-type (top inlet) receiver. The length of horizontal liquid drain lines to the receiver should be minimized, with no traps allowed. Equalization between the shells is achieved by keeping liquid velocity in the drain line less than 0.5 m/s. The drain line can be sized from Table 2.

Evaporative Condensers Evaporative condensers are selected based on the wet-bulb temperature in which they operate. The 1% design wet bulb is that temperature that will be equaled or exceeded 1% of the months of June through September, or 29.3 h. Thus, for the majority of industrial plants that operate at least at part load all year, the wet-bulb temperature is below design 99.6% of the operating time. The resultant condensing pressure only equals or exceeds the design condition 0.4% of the time if the design wet-bulb temperature and

Fig. 10 Parallel Condensers with Top Inlet Receiver peak design refrigeration load occur coincidentally. This peak condition is more a function of how the load is calculated, what load diversity factor exists or is used in the calculation, and what safety factor is used in the calculations, than of the size of the condenser. Location. If an evaporative condenser is located with insufficient space for air movement, the effect is the same as that imposed by an inlet damper, and the fan may not deliver enough air. In addition, evaporative condenser discharge air may recirculate, which adds to the problem. The high inlet velocity causes a

Ammonia Refrigeration Systems low-pressure region to develop around the fan inlet, inducing flow of discharge air into that region. If the obstruction is from a second condenser, the problem can be even more severe because discharge air from the second condenser flows into the air intake of the first. Prevailing winds can also contribute to recirculation. In many areas, winds shift with the seasons; wind direction during the peak high-humidity season is the most important consideration. The tops of condensers should always be higher than any adjacent structure to eliminate downdrafts that might induce recirculation. Where this is impractical, discharge hoods can be used to discharge air far enough away from the fan intakes to avoid recirculation. However, the additional static pressure imposed by a discharge hood must be added to the fan system. Fan speed can be increased slightly to obtain proper air volume. Installation. A single evaporative condenser used with a through-type (top inlet) receiver can be connected as shown in Figure 11. The receiver must always be at a lower pressure than the condensing pressure. Proper design ensures that the receiver is cooler than the condensing temperature. In areas with ambient temperatures below 0°C, water in the evaporative condenser drain pan and water circuit must be kept from freezing at light plant loads. When the temperature is at freezing, the evaporative condenser can operate as a dry-coil unit, and the water pump(s) and piping can be drained and secured for the season. Another method of keeping water from freezing is to place the water tank inside and install it as illustrated in Figure 12. When outdoor temperature drops, the condensing pressure drops, and a pressure switch with its sensing element in the discharge pressure line stops the water pump; the water is then drained into the tank. An alternative is to use a thermostat that senses water or outdoor ambient temperature and stops the pump at low temperatures. Exposed piping and any trapped water headers in the evaporative condenser should be drained into the indoor water tank. Air volume capacity control methods include inlet, outlet, or bypass dampers; two-speed or VFD fan motors; or fan cycling in response to pressure controls. Liquid Traps. Because all evaporative condensers have substantial pressure drop in the ammonia circuit, liquid traps are needed at

2.7 the outlets when two or more condensers or condenser coils are installed (see Figure 13). Also, an equalizer line is necessary to maintain stable pressure in the receiver to ensure free drainage from condensers. For example, assume a 10 kPa pressure drop in the operating condenser in Figure 13, which produces a lower pressure (1290 kPa) at its outlet compared to the idle condenser (1300 kPa) and the receiver (1300 kPa). The trap creates a liquid seal so that a liquid height h of 1700 mm (equivalent to 10 kPa) builds up in the vertical drop leg and not in the condenser coil.

Fig. 12 Evaporative Condenser with Inside Water Tank

Fig. 11 Single Evaporative Condenser with Top Inlet Receiver

Fig. 13 Two Evaporative Condensers with Trapped Piping to Receiver

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 14 Method of Reducing Condenser Outlet Sizes

Fig. 16 Piping for Parallel Condensers with Surge-Type Receiver conditions and is based on the maximum condensing pressure drop of the coil. If service valves are installed at the coil inlets and/or outlets, the pressure drops imposed by these valves must be accounted for by increasing the minimum 1500 mm drop-leg height by an amount equal to the valve pressure drop in height of liquid refrigerant (Figure 14). Figures 15, 16, and 17 illustrate various piping arrangements for evaporative condensers.

EVAPORATORS

Fig. 15 Piping for Shell-and-Tube and Evaporative Condensers with Top Inlet Receiver The trap must have enough height above the vertical liquid leg to accommodate a liquid height equal to the maximum pressure drop encountered in the condenser. The example illustrates the extreme case of one unit on and one off; however, the same phenomenon occurs to a lesser degree with two condensers of differing pressure drops when both are in full operation. Substantial differences in pressure drop can also occur between two different brands of the same size condenser or even different models produced by the same manufacturer. The minimum recommended height of the vertical leg is 1500 mm for ammonia. This vertical dimension h is shown in all evaporative condenser piping diagrams. This height is satisfactory for operation within reasonable ranges around normal design

Several types of evaporators are used in ammonia refrigeration systems. Fan-coil, direct-expansion evaporators can be used, but they are not generally recommended unless the suction temperature is –18°C or higher. This is due in part to the relative inefficiency of the direct-expansion coil, but more importantly, the low mass flow rate of ammonia is difficult to feed uniformly as a liquid to the coil. Instead, ammonia fan-coil units designed for recirculation (overfeed) systems are preferred. Typically, in this type of system, high-pressure ammonia from the system high stage flashes into a large vessel at the evaporator pressure, from which it is pumped to the evaporators at an overfeed rate of 2.5:1 to 4:1. This type of system is standard and very efficient. See Chapter 4 for more details. Flooded shell-and-tube evaporators are often used in ammonia systems in which indirect or secondary cooling fluids such as water, brine, or glycol must be cooled. Some problems that can become more acute at low temperatures include changes in lubricant transport properties, loss of capacity caused by static pressure from the depth of the pool of liquid refrigerant in the evaporator, deterioration of refrigerant boiling heat transfer coefficients caused by lubricant logging, and higher specific volumes for the vapor. The effect of pressure losses in the evaporator and suction piping is more acute in low-temperature systems because of the large change in saturation temperatures and specific volume in relation to

Ammonia Refrigeration Systems

2.9

Fig. 18 Piping for Thermostatic Expansion Valve Application for Automatic Defrost on Unit Cooler

Fig. 17

Piping for Parallel Condensers with Top Inlet Receiver

pressure changes at these conditions. Systems that operate near or below zero gage pressure are particularly affected by pressure loss. The depth of the pool of boiling refrigerant in a flooded evaporator exerts a liquid pressure on the lower part of the heat transfer surface. Therefore, the saturation temperature at this surface is higher than that in the suction line, which is not affected by the liquid pressure. This temperature gradient must be considered when designing the evaporator. Spray shell-and-tube evaporators, though not commonly used, offer certain advantages. In this design, the evaporator’s liquid depth penalty can be eliminated because the pool of liquid is below the heat transfer surface. A refrigerant pump sprays liquid over the surface. Pump energy is an additional heat load to the system, and more refrigerant must be used to provide the net positive suction pressure required by the pump. The pump is also an additional item that must be maintained. This evaporator design also reduces the refrigerant charge requirement compared to a flooded design (see Chapter 4).

EVAPORATOR PIPING Proper evaporator piping and control are necessary to keep the cooled space at the desired temperature and also to adequately protect the compressor from surges of liquid ammonia out of the evaporator. The evaporators illustrated in this section show some methods used to accomplish these objectives. In some cases, combinations of details on several illustrations have been used. When using hot gas or electric heat for defrosting, the drain pan and drain line must be heated to prevent the condensate from refreezing. With hot gas, a heating coil is embedded in the drain pan. The hot gas flows first through this coil and then into the evaporator coil.

With electric heat, an electric heating coil is used under the drain pan. Wraparound or internal electric heating cables are used on the condensate drain line when the room temperature is below 0°C. Figure 18 illustrates a thermostatic expansion valve on a unit cooler using hot gas for automatic defrosting. Because this is an automatic defrosting arrangement, hot gas must always be available at the hot-gas solenoid valve near the unit. The system must contain multiple evaporators so the compressor is running when the evaporator to be defrosted is shut down. The hot-gas header must be kept in a space where ammonia does not condense in the pipe. Otherwise, the coil receives liquid ammonia at the start of defrosting and is unable to take full advantage of the latent heat of hot-gas condensation entering the coil. This can also lead to severe hydraulic shock loads. If the header must be in a cold space, the hot-gas main must be insulated and a high-pressure float drainer installed to remove any accumulated condensate. The liquid- and suction-line solenoid valves are open during normal operation only and are closed during the defrost cycle. When defrost starts, the hot-gas solenoid valve is opened. Refer to IIAR Bulletin 116 for information on possible hydraulic shock when the hot-gas defrost valve is opened after a defrost. A defrost pressure regulator maintains a gage pressure of about 480 to 550 kPa in the coil.

Unit Cooler: Flooded Operation Figure 19 shows a flooded evaporator with a close-coupled lowpressure vessel for feeding ammonia into the coil and automatic water defrost. The lower float switch on the float column at the vessel controls opening and closing of the liquid-line solenoid valve, regulating ammonia feed into the unit to maintain a liquid level. The hand expansion valve downstream of the solenoid valve should be adjusted so that it does not feed ammonia into the vessel more quickly than the vessel can accommodate while raising the suction pressure of gas from the vessel no more than 7 to 14 kPa. The static height of liquid in the vessel should be sufficient to flood the coil with liquid under normal loads. The higher float switch is to signal a high level of liquid in the vessel. It should be wired into an alarm circuit or possibly a compressor shutdown circuit if there is no other compressor protection. The float switches and/or columns should be insulated. With flooded coils having horizontal headers, distribution between the multiple circuits is accomplished without distributing orifices.

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 20 Fig. 19

Arrangement for Automatic Defrost of Air Blower with Flooded Coil

A combination evaporator pressure regulator and stop valve is used in the suction line from the vessel. During operation, the regulator maintains a nearly constant back pressure in the vessel. A solenoid coil in the regulator mechanism closes it during the defrost cycle. The liquid solenoid valve should also be closed at this time. One of the best means of controlling room temperature is a room thermostat that controls the effective setting of the evaporator pressure regulator. A spring-loaded relief valve is used around the suction pressure regulator and is set so that the vessel is kept below 860 kPa (gage). Other suction line pressure control arrangements, such as a dual pressure regulator, can be used to eliminate the extra piping of the relief valve. A solenoid valve unaffected by downstream pressure is used in the water line to the defrost header. The defrost header is constructed so that it drains at the end of the defrost cycle and the downstream side of the solenoid valve drains through a fixed orifice. Unless the room is kept above 0°C, the drain line from the unit should be wrapped with a heater cable or provided with another heat source and then insulated to prevent defrost water from refreezing in the line. Water line length in the space leading up to the header and the length of the drain line in the cooled space should be minimized. A flapper or pipe trap on the end of the drain line prevents warm air from flowing up the drain pipe and into the unit. An air outlet damper may be closed during defrosting to prevent thermal circulation of air through the unit, which affects the temperature of the cooled space. The fan is stopped during defrost. This type of defrosting requires a drain pan float switch for safety control. If the drain pan fills with water, the switch overrides the time clock to stop flow into the unit by closing the water solenoid valve. There should be a 5 min delay at the end of the water spray part of the defrosting cycle so water can drain from the coil and pan. This limits ice build-up in the drain pan and on the coils after the cycle is completed. On completion of the cycle, the low-pressure vessel may be at about 517 kPa (gage). When the unit is opened to the much-lowerpressure suction main, some liquid surges out into the main; therefore, it may be necessary to gradually bleed off this pressure before fully opening the suction valve, to prevent thermal shock. Generally, a suction trap in the engine room removes this liquid before the gas stream enters the compressors.

Arrangement for Horizontal Liquid Cooler and High-Side Float

The type of refrigerant control shown in Figure 19 can be used on brine spray units where brine is sprayed over the coil at all times to pick up the condensed water vapor from the airstream. The brine is reconcentrated continually to remove water absorbed from the airstream.

High-Side Float Control When a system has only one evaporator, a high-pressure float control can be used to keep the condenser drained and to provide a liquid seal between the high and low sides. Figure 20 illustrates a brine or water cooler with this type of control. The high-side float should be located near the evaporator to avoid insulating the liquid line. The amount of ammonia in this type of system is critical because the charge must be limited so that liquid will not surge into the suction line under the highest loading in the evaporator. Some type of suction trap should be used. One method is to place a horizontal shell above the cooler, with suction gas piped into the bottom and out the top. The reduction of gas velocity in this shell causes liquid to separate from the gas and drop back into the chiller. Coolers should include a liquid indicator. A reflex glass lens with a large liquid chamber and vapor connections for boiling liquids and a plastic frost shield to determine the actual level should be used. A refrigeration thermostat measuring chilled-fluid temperature as it exits the cooler should be wired into the compressor starting circuit to prevent freezing. A flow switch or differential pressure switch should prove flow before the compressor starts. The fluid to be cooled should be piped into the lower portion of the tube bundle and out of the top portion.

Low-Side Float Control For multiple evaporator systems, low-side float valves are used to control the refrigerant level in flooded evaporators. The lowpressure float in Figure 21 has an equalizer line from the top of the float chamber to the space above the tube bundle and an equalizer line out of the lower side of the float chamber to the lower side of the tube bundle. For positive shutoff of liquid feed when the system stops, a solenoid valve in the liquid line is wired so that it is only energized when the brine or water pump motor is operating and the compressor is running. A reflex glass lens with large liquid chamber and vapor connections for boiling liquids should be used with a plastic frost shield to

Ammonia Refrigeration Systems

Fig. 21

2.11

Piping for Evaporator and Low-Side Float with Horizontal Liquid Cooler

determine the actual level, and with front extensions as required. These chambers or columns should be insulated to prevent false levels caused by heat transfer from the surrounding environment. Usually a high-level float switch is installed above the operating level of the float to shut the liquid solenoid valve if the float should overfeed.

VESSELS High-Pressure Receivers. Industrial systems generally incorporate a central high-pressure refrigerant receiver, which serves as the primary refrigerant storage location in the system. It handles refrigerant volume variations between the condenser and the system’s low side during operation and pumpdowns for repairs or defrost. Ideally, the receiver should be large enough to hold the entire system charge, but this is not generally economical. The system should be analyzed to determine the optimum receiver size. Receivers are commonly equalized to the condenser inlet and operate at the same pressure as the condenser. In some systems, the receiver is operated at a pressure between the condensing pressure and the highest suction pressure to allow for variations in condensing pressure without affecting the system’s feed pressure. This type is commonly referred to as a controlled-pressure receiver (CPR). Liquid from the condenser is metered through a high-side control as it is condensed. CPR pressure is maintained with a back-pressure regulator vented to an intermediate pressure point. Winter or low-load operating conditions may require a downstream pressure regulator to maintain a minimum pressure. If additional receiver capacity is needed for normal operation, use extreme caution in the design. Designers usually remove the inadequate receiver and replace it with a larger one rather than install an additional receiver in parallel. This procedure is best because even slight differences in piping pressure or temperature can cause the refrigerant to migrate to one receiver and not to the other. Smaller auxiliary receivers can be incorporated to serve as sources of high-pressure liquid for compressor injection or thermosiphon, lubricant cooling, high-temperature evaporators, and so forth. Intercoolers (Gas and Liquid). An intercooler (subcooler/ desuperheater) is the intermediate vessel between the high and low stages in a multistage system. One purpose is to cool discharge gas of the low-stage compressor to prevent overheating the high-stage compressor. This can be done by bubbling discharge gas from the low-stage compressor through a bath of liquid refrigerant or by mixing liquid normally entering the intermediate vessel with the discharge gas as it enters above the liquid level. Heat removed from the discharge gas is absorbed by evaporating part of the liquid and

Fig. 22 Intercooler eventually passes through the high-stage compressor to the condenser. Distributing the discharge gas below a level of liquid refrigerant separates out any lubricant carryover from the low-stage compressor. If liquid in the intercooler is to be used for other purposes, such as liquid makeup or feed to the low stage, periodic lubricant removal is imperative. Another purpose of the intercooler is to lower the liquid temperature used in the low stage of a two-stage system. Lowering refrigerant temperature in the intercooler with high-stage compressors increases the refrigeration effect and reduces the low-stage compressor’s required displacement, thus reducing its operating cost. Intercoolers for two-stage compression systems can be shelland-coil or flash. Figure 22 depicts a shell-and-coil intercooler incorporating an internal pipe coil for subcooling high-pressure liquid before it is fed to the low stage of the system. Typically, the coil subcools liquid to within 6 K of the intermediate temperature. Vertical shell-and-coil intercoolers perform well in many applications using ammonia refrigerant systems. Horizontal designs are possible but usually not practical. The vessel must be sized properly to separate liquid from vapor that is returning to the high-stage compressor. The superheated gas inlet pipe should extend below the liquid level and have perforations or slots to distribute the gas evenly in small bubbles. Adding a perforated baffle across the area of the vessel slightly below the liquid level protects against violent surging. Always use a float switch that shuts down the high-stage compressor when the liquid level gets too high. A means of maintaining a liquid level for the subcooling coil and low-stage compressor desuperheating is necessary if no high-stage evaporator overfeed liquid is present. Electronic level controls (see Figure 29) can simplify the use of multiple float switches and float valves to maintain the various levels required. The flash intercooler is similar in design to the shell-and-coil intercooler, except for the coil. The high-pressure liquid is flashcooled to the intermediate temperature. Use caution in selecting a flash intercooler because all the high-pressure liquid is flashed to intermediate pressure. Though colder than that of the shell-and-coil intercooler, liquid in the flash intercooler is not subcooled and is susceptible to flashing from system pressure drop. Two-phase liquid feed to control valves may cause premature failure because of the wire-drawing effect of the liquid/vapor mixture.

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 23

Arrangement for Compound System with Vertical Intercooler and Suction Trap

Figure 23 shows a vertical shell-and-coil intercooler as piped into the system. The liquid level is maintained in the intercooler by a float that controls the solenoid valve feeding liquid into the shell side of the intercooler. Gas from the first-stage compressor enters the lower section of the intercooler, is distributed by a perforated plate, and is then cooled to the saturation temperature corresponding to intermediate pressure. When sizing any intercooler, the designer must consider (1) lowstage compressor capacity; (2) vapor desuperheating, liquid makeup requirements for the subcooling coil load, or vapor cooling load associated with the flash intercooler; and (3) any high-stage side loading. The volume required for normal liquid levels, liquid surging from high-stage evaporators, feed valve malfunctions, and liquid/vapor must also be analyzed. Necessary accessories are the liquid level control device and high-level float switch. Though not absolutely necessary, an auxiliary oil pot should also be considered. Suction Accumulator. A suction accumulator (also known as a knockout drum, suction trap, etc.) prevents liquid from entering the suction of the compressor, whether on the high or low stage of the system. Both vertical and horizontal vessels can be incorporated. Baffling and mist eliminator pads can enhance liquid separation. Suction accumulators, especially those not intentionally maintaining a level of liquid, should have a way to remove any build-up of ammonia liquid. Gas boil-out coils or electric heating elements are costly and inefficient. Although it is one of the more common and simplest means of liquid removal, a liquid boil-out coil (Figure 24) has some drawbacks. Generally, warm liquid flowing through the coil is the heat source for liquid being boiled off. Liquid transfer pumps, gaspowered transfer systems, or basic pressure differentials are a more positive means of removing the liquid (Figures 25 and 26). Accessories should include a high-level float switch for compressor protection along with additional pump or transfer system controls. Vertical Suction Trap and Pump. Figure 27 shows the piping of a vertical suction trap that uses a high-pressure ammonia pump to transfer liquid from the system’s low-pressure side to the highpressure receiver. Float switches piped on a float column on the side

Fig. 24 Suction Accumulator with Warm Liquid Coil

of the trap can start and stop the liquid ammonia pump, sound an alarm in case of excess liquid, and sometimes stop the compressors. When the liquid level in the suction trap reaches the setting of the middle float switch, the liquid ammonia pump starts and reduces the liquid level to the setting of the lower float switch, which stops the liquid ammonia pump. A check valve in the discharge line of the ammonia pump prevents gas and liquid from flowing backward through the pump when it is not in operation. Depending on the type of check valve used, some installations have two valves in a series as an extra precaution against pump backspin. Compressor controls adequately designed for starting, stopping, and capacity reduction result in minimal agitation, which helps separate vapor and liquid in the suction trap. Increasing compressor

Ammonia Refrigeration Systems

Fig. 25

2.13

Equalized Pressure Pump Transfer System

Fig. 27 Piping for Vertical Suction Trap and High-Pressure Pump

Fig. 26 Gravity Transfer System capacity slowly and in small increments reduces liquid boiling in the trap, which is caused by the refrigeration load of cooling the refrigerant and metal mass of the trap. If another compressor is started when plant suction pressure increases, it should be brought on line slowly to prevent a sudden pressure change in the suction trap. A high level of liquid in a suction trap should activate an alarm or stop the compressors. Although eliminating the cause is the most effective way to reduce a high level of excess surging liquid, a more immediate solution is to stop part of the compression system and raise plant suction pressure slightly. Continuing high levels indicate insufficient pump capacity or suction trap volume. Liquid Level Indicators. Liquid level can be indicated by visual indicators, electronic sensors, or a combination of the two. Visual indicators include individual circular reflex level indicators (bull’s-eyes) mounted on a pipe column or stand-alone linear reflex glass assemblies (Figure 28). For operation at temperatures below the frost point, transparent plastic frost shields covering the reflex surfaces are necessary. Also, the pipe column must be insulated, especially when control devices are attached to prevent false level readings caused by heat influx. Electronic level sensors can continuously monitor liquid level. Digital or graphic displays of liquid level can be locally or remotely monitored (Figure 29). Level indicators should have adequate isolation valves, which should incorporate stop check or excess-flow valves for isolation and safety.

Fig. 28

Gage Glass Assembly for Ammonia

Purge Units. A noncondensable gas separator (purge unit) is useful in most plants, especially when suction pressure is below atmospheric pressure. Purge units on ammonia systems are piped to carry noncondensables (air) from the receiver and condenser to the purger, as shown in Figure 30. Suction from the coil should be taken to one of the low-temperature suction vessel inlet mains. Ammonia vapor and noncondensable gas are drawn into the purger, and the ammonia condenses on the cold surface, sorting out the noncondensables. When the drum fills with air and other noncondensables, a level control in the purger opens and allows them to be released. Depending on operating conditions, a trace of ammonia may remain in the noncondensable gases. The noncondensable gases are diverted to a water bottle (generally with running water) to diffuse the pungent odor of the ammonia. Ammonia systems, which are inherently large,

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 29

Electronic Liquid Level Control

2. Liquid lines 50 to 150 mm shall be not less than Schedule 40 carbon steel pipe. 3. Liquid lines 200 to 300 mm shall be not less than Schedule 20 carbon steel pipe. 4. Vapor lines 150 mm and smaller shall be not less than Schedule 40 carbon steel pipe. 5. Vapor lines 200 to 300 mm shall be not less than Schedule 20 carbon steel pipe. 6. Vapor lines 350 mm and larger shall be not less than Schedule 10 carbon steel pipe. 7. All threaded pipe shall be Schedule 80. 8. Carbon steel pipe shall be ASTM Standard A53 Grade A or B, Type E (electric resistance welded) or Type S (seamless); or ASTM Standard A106 (seamless), except where temperaturepressure criteria mandate a higher specification material. Standard A53 Type F is not permitted for ammonia piping.

Fittings Couplings, elbows, and tees for threaded pipe are for a minimum of 21 MPa design pressure and constructed of forged steel. Fittings for welded pipe should match the type of pipe used (i.e., standard fittings for standard pipe and extra-heavy fittings for extra-heavy pipe). Tongue-and-groove or ANSI flanges should be used in ammonia piping. Welded flanges for low-side piping can have a minimum 1 MPa design pressure rating. On systems located in high ambients, low-side piping and vessels should be designed for 1.4 to 1.6 MPa. The high side should be 1.7 MPa if the system uses watercooled or evaporative cooled condensing. Use 2.1 MPa minimum for air-cooled designs.

Pipe Joints

Fig. 30 Noncondensable Gas Purger Unit have multiple points where noncondensables can collect. Purge units that can automatically sequence through the various points and remove noncondensables are available. Ammonia’s affinity for water poses another system efficiency concern. The presence of water increases the refrigerant temperature above the saturated pressure. The increased temperature requires lower operating pressures to maintain the same refrigerant temperature. Unlike noncondensable gases, which collect in the system’s high side and result in higher condensing pressures, the presence of water is less obvious. Water collects in the liquid phase and forms an aqua/ammonia solution. Short of a complete system charge removal, distillers (temporary or permanent) can be incorporated. Automatic noncondensable and water removal units can provide continual monitoring of the system impurities.

PIPING Local codes or ordinances governing ammonia mains should be followed, in addition to the recommendations here.

Recommended Material Because copper and copper-bearing materials are attacked by ammonia, they are not used in ammonia piping systems. Steel or stainless steel piping, fittings, and valves of the proper pressure rating are suitable for ammonia gas and liquid. Ammonia piping should conform to ASME Standard B31.5, and to IIAR Standard 2, which states the following: 1. Liquid lines 40 mm and smaller shall be not less than Schedule 80 carbon steel pipe.

Joints between lengths of pipe or between pipe and fittings can be threaded if the pipe size is 32 mm or smaller. Pipe 40 mm or larger should be welded. An all-welded piping system is superior. Threaded Joints. Many sealants and compounds are available for sealing threaded joints. The manufacturer’s instructions cover compatibility and application method. Do not use excessive amounts or apply on female threads because any excess can contaminate the system. Welded Joints. Pipe should be cut and beveled before welding. Use pipe alignment guides and provide a proper gap between pipe ends so that a full-penetration weld is obtained. The weld should be made by a qualified welder, using proper procedures such as the Welding Procedure Specifications, prepared by the National Certified Pipe Welding Bureau (NCPWB). Gasketed Joints. A compatible fiber gasket should be used with flanges. Before tightening flange bolts to valves, controls, or flange unions, properly align pipe and bolt holes. When flanges are used to straighten pipe, they put stress on adjacent valves, compressors, and controls, causing the operating mechanism to bind. To prevent leaks, flange bolts are drawn up evenly when connecting the flanges. Flanges at compressors and other system components must not move or indicate stress when all bolts are loosened. Union Joints. Steel (21 MPa) ground joint unions are used for gage and pressure control lines with screwed valves and for joints up to 20 mm. When tightening this type of joint, the two pipes must be axially aligned. To be effective, the two parts of the union must match perfectly. Ground joint unions should be avoided if at all possible.

Pipe Location Piping should be at least 2.3 m above the floor. Locate pipes carefully in relation to other piping and structural members, especially when lines are to be insulated. The distance between insulated lines should be at least three times the thickness of the insulation for screwed

Ammonia Refrigeration Systems

2.15

Table 1 Suction Line Capacities in Kilowatts for Ammonia with Pressure Drops of 0.005 and 0.01 K/m Equivalent Saturated Suction Temperature, °C –50

–40

–30

Steel Nominal Line Size, mm

t = 0.005 K/m p = 12.1 Pa/m

t = 0.01 K/m p = 24.2 Pa/m

t = 0.005 K/m p = 19.2 Pa/m

t = 0.01 K/m p = 38.4 Pa/m

t = 0.005 K/m p = 29.1 Pa/m

t = 0.01 K/m p = 58.2 Pa/m

10 15 20 25 32 40 50 65 80 100 125 150 200 250 300

0.19 0.37 0.80 1.55 2.39 3.68 9.74 15.67 28.08 57.95 105.71 172.28 356.67 649.99 1045.27

0.29 0.55 1.18 2.28 3.51 5.41 14.22 22.83 40.81 84.10 153.05 248.91 514.55 937.58 1504.96

0.35 0.65 1.41 2.72 4.43 6.85 16.89 27.13 48.36 99.50 181.16 294.74 609.20 1107.64 1777.96

0.51 0.97 2.08 3.97 6.47 9.94 24.50 39.27 69.99 143.84 261.22 424.51 874.62 1589.51 2550.49

0.58 1.09 2.34 4.48 7.66 11.77 27.57 44.17 78.68 161.77 293.12 476.47 981.85 1782.31 2859.98

0.85 1.60 3.41 6.51 11.14 17.08 39.82 63.77 113.30 232.26 420.83 683.18 1402.03 2545.46 4081.54

Saturated Suction Temperature, °C 20

5

Steel Nominal Line Size, mm

t = 0.005 K/m p = 42.2 Pa/m

t = 0.01 K/m p = 84.4 Pa/m

t = 0.005 K/m p = 69.2 Pa/m

10 15 20 25 32 40 50 65 80 100 125 150 200 250 300

0.91 1.72 3.66 6.98 12.47 19.08 42.72 68.42 121.52 249.45 452.08 733.59 1506.11 2731.90 4378.87

1.33 2.50 5.31 10.10 18.03 27.48 61.51 98.23 174.28 356.87 646.25 1046.77 2149.60 3895.57 6237.23

1.66 3.11 6.61 12.58 19.22 29.45 76.29 122.06 216.15 442.76 800.19 1296.07 2662.02 4818.22 7714.93

+5 t = 0.01 K/m p = 138.3 Pa/m 2.41 4.50 9.53 18.09 28.67 42.27 109.28 174.30 308.91 631.24 1139.74 1846.63 3784.58 6851.91 10 973.55

t = 0.005 K/m p = 92.6 Pa/m 2.37 4.42 9.38 17.79 28.32 43.22 107.61 171.62 304.12 621.94 1124.47 1819.59 3735.65 6759.98 10 810.65

t = 0.01 K/m p = 185.3 Pa/m 3.42 6.37 13.46 25.48 36.02 54.88 153.66 245.00 433.79 885.81 1598.31 2590.21 5303.12 9589.56 15 360.20

Note: Capacities are in kilowatts of refrigeration resulting in a line friction loss per unit equivalent pipe length (p in Pa/m), with corresponding change in saturation temperature per unit length (t in K/m).

fittings, and four times for flange fittings. The space between the pipe and adjacent surfaces should be three-fourths of these amounts. Hangers located close to the vertical risers to and from compressors keep the piping weight off the compressor. Pipe hangers should be placed no more than 2.4 to 3 m apart, depending on pipe size, and within 0.6 m of a change in direction of the piping. Hangers should be designed to bear on the outside of insulated lines. Sheet metal sleeves on the lower half of the insulation are usually sufficient. Where piping penetrates a wall, a sleeve should be installed, and where the pipe penetrating the wall is insulated, it must be adequately sealed. Piping to and from compressors and to other components must provide for expansion and contraction. Sufficient flange or union joints should be located in the piping so components can be assembled easily during installation and also disassembled for servicing.

Pipe Sizing Table 1 presents practical suction line sizing data based on 0.005 K and 0.01 K differential pressure drop equivalent per metre total equivalent length of pipe, assuming no liquid in the suction line. For data on equivalent lengths of valves and fittings, refer to Tables 10, 11, and 12 in Chapter 1. Table 2 lists data for sizing

suction and discharge lines at 0.02 K differential pressure drop equivalent per metre equivalent length of pipe, and for sizing liquid lines at 0.5 m/s. Charts prepared by Wile (1977) present pressure drops in saturation temperature equivalents. For a complete discussion of the basis of these line sizing charts, see Timm (1991). Table 3 presents line sizing information for pumped liquid lines, high-pressure liquid lines, hot-gas defrost lines, equalizing lines, and thermosiphon lubricant cooling ammonia lines.

CONTROLS Refrigerant flow controls are discussed in Chapter 11. The following precautions are necessary in the application of certain controls in low-temperature systems.

Liquid Feed Control Many controls available for single-stage, high-temperature systems may be used with some discretion on low-temperature systems. If the liquid level is controlled by a low-side float valve (with the float in the chamber where the level is controlled), low pressure and temperature have no appreciable effect on operation. External float chambers, however, must be thoroughly insulated to prevent heat influx that might cause boiling and an unstable level, affecting

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2014 ASHRAE Handbook—Refrigeration (SI)

Table 2 Suction, Discharge Line, and Liquid Capacities in Kilowatts for Ammonia (Single- or High-Stage Applications) Discharge Lines t = 0.02 K/m, p = 684.0 Pa/m

Suction Lines (t = 0.02 K/m) Steel Saturated Suction Temperature, °C Nominal Line Size, –40 –30 –20 –5 +5 mm p = 76.9 p = 116.3 p = 168.8 p = 276.6 p = 370.5

Saturated Suction Temp., °C –40

–20

+5

Liquid Lines Steel Nominal Line Size, mm

Velocity = 0.5 m/s

p = 450.0

10 15 20

0.8 1.4 3.0

1.2 2.3 4.9

1.9 3.6 7.7

3.5 6.5 13.7

4.9 9.1 19.3

8.0 14.9 31.4

8.3 15.3 32.3

8.5 15.7 33.2

10 15 20

3.9 63.2 110.9

63.8 118.4 250.2

25 32 40

5.8 9.5 14.4

9.4 16.16 24.60

14.6 25.7 39.4

25.9 46.4 60.4

36.4 57.6 88.2

59.4 96.2 146.0

61.0 107.0 163.8

62.6 98.9 151.4

25 32 40

179.4 311.0 423.4

473.4 978.0 1469.4

50 65 80

35.4 56.7 101.0

57.2 91.6 162.4

88.1 140.6 249.0

155.7 248.6 439.8

218.6 348.9 616.9

355.2 565.9 1001.9

364.9 581.4 1029.3

374.7 597.0 1056.9

50 65 80

697.8 994.8 1536.3

2840.5 4524.8 8008.8

100 125 150 200

206.9 375.2 608.7 1252.3

332.6 601.8 975.6 2003.3

509.2 902.6 1491.4 3056.0

897.8 1622.0 2625.4 5382.5

1258.6 2271.4 3672.5 7530.4

2042.2 3682.1 5954.2 12 195.3

2098.2 3783.0 6117.4 12 529.7

2154.3 3884.2 6281.0 12 864.8

— — — —

— — — —

— — — —

250 300

2271.0 3640.5

3625.9 5813.5

5539.9 8873.4

9733.7 15568.9

13619.6 21787.1

22 028.2 35 239.7

22 632.2 36 206.0

23 237.5 37 174.3

— —

— —

— —

Notes: 1. Table capacities are in kilowatts of refrigeration.

4. Values are based on 30°C condensing temperature. Multiply table capacities by the following factors for other condensing temperatures:

p = pressure drop due to line friction, Pa/m t = corresponding change in saturation temperature, K/m 2. Line capacity for other saturation temperatures t and equivalent lengths Le

Condensing Temperature, °C 20 30 40 50

Table L Actual t 0.55 Line capacity = Table capacity  ----------------------e-  -----------------------  Actual L e Table t  3. Saturation temperature t for other capacities and equivalent lengths Le Actual L Actual capacity 1.8 t = Table t  -----------------------e  -------------------------------------  Table L e   Table capacity 

Suction Lines 1.04 1.00 0.96 0.91

Discharge Lines 0.86 1.00 1.24 1.43

5. Liquid line capacities based on 5°C suction.

Table 3 Liquid Ammonia Line Capacities in Kilowatts Nominal Size, mm 15 20 25 32 40 50 65 80 100 125 150 200

3:1

4:1

5:1

High-Pressure Liquid at 21 kPaa

35 77 151 329 513 1175 1875 2700 4800 — — —

26 58 114 246 387 879 1407 2026 3600 — — —

21 46 92 197 308 703 1125 1620 2880 — — —

106 243 472 1007 1544 3573 5683 10 150 — — — —

Pumped Liquid Overfeed Ratio

Source: Wile (1977). for hot-gas branch lines under 30 m with minimum inlet pressure of 724 kPa (gage), defrost pressure of 483 kPa (gage), and –29°C evaporators designed for a 5.6 K temperature differential

aRating

the float response. Equalizing lines to external float chambers, particularly the upper line, must be sized generously so that liquid can reach the float chamber, and gas resulting from any evaporation may be returned to the vessel without appreciable pressure loss. The superheat-controlled (thermostatic) expansion valve is generally used in direct-expansion evaporators. This valve operates on the difference between bulb pressure, which is responsive to suction temperature, and pressure below the diaphragm, which is the actual suction pressure.

Hot-Gas Defrosta

Equalizer High Sideb

32 56 99 106 176 324 570 1154 2089 3411 —

176 352 528 791 1055 1759 3517 7034 — — —

b Line

Thermosiphon Lubricant Cooling Lines Gravity Flowc Supply

Return

Vent

59 138 249 385 663 1041 1504 2600

35 88 155 255 413 649 938 1622

60 106 187 323 586 1062 1869 3400

sizes based on experience using total system evaporator kilowatts. Frick Co. (1995). Values for line sizes above 100 mm are extrapolated.

c From

The thermostatic expansion valve is designed to maintain a preset superheat in suction gas. Although the pressure-sensing part of the system responds almost immediately to a change in conditions, the temperature-sensing bulb must overcome thermal inertia before its effect is felt on the power element of the valve. Thus, when compressor capacity decreases suddenly, the expansion valve may overfeed before the bulb senses the presence of liquid in the suction line and reduces the feed. Therefore, a suction accumulator should be installed on direct-expansion low-temperature systems with multiple expansion valves.

Ammonia Refrigeration Systems Controlling Load During Pulldown System transients during pulldown can be managed by controlling compressor capacity. Proper load control reduces compressor capacity so that energy requirements stay within the motor and condenser capacities. On larger systems using screw compressors, a current-sensing device reads motor amperage and adjusts the capacity control device appropriately. Cylinders on reciprocating compressors can be unloaded for similar control. Alternatively, a downstream, outlet, or crankcase pressure regulator can be installed in the suction line to throttle suction flow if the pressure exceeds a preset limit. This regulator limits the compressor’s suction pressure during pulldown. The disadvantage of this device is the extra pressure drop it causes when the system is at the desired operating conditions. To overcome some of this, the designer can use external forces to drive the valve, causing it to be held fully open when the pressure is below the maximum allowable. Systems using downstream pressure regulators and compressor unloading must be carefully designed so that the two controls complement each other.

Operation at Varying Loads and Temperatures Compressor and evaporator capacity controls are similar for multi- and single-stage systems. Control methods include compressor capacity control, hot-gas bypass, or evaporator pressure regulators. Low pressure can affect control systems by significantly increasing the specific volume of the refrigerant gas and the pressure drop. A small pressure reduction can cause a large percentage capacity reduction. System load usually cannot be reduced to near zero, because this results in little or no flow of gas through the compressor and consequent overheating. Additionally, high pressure ratios are detrimental to the compressor if it is required to run at very low loads. If the compressor cannot be allowed to cycle off during low load, an acceptable alternative is a hot-gas bypass. High-pressure gas is fed to the low-pressure side of the system through a downstream pressure regulator. The gas should be desuperheated by injecting it at a point in the system where it is in contact with expanding liquid, such as immediately downstream of the liquid feed to the evaporator. Otherwise, extremely high compressor discharge temperatures can result. The artificial load supplied by high-pressure gas can fill the gap between the actual load and the lowest stable compressor operating capacity. Figure 31 shows such an arrangement.

Electronic Control Microprocessor- and computer-based control systems are the norm for control systems on individual compressors as well as for entire system control. Almost all screw compressors use microprocessor control systems to monitor all safety functions and operating conditions. These machines are frequently linked together with a programmable controller or computer for sequencing multiple compressors so that they load and unload in response to system fluctuations in the most economical manner. Programmable controllers are also used to replace multiple defrost time clocks on larger systems for more

2.17 accurate and economical defrosting. Communications and data logging allow systems to operate at optimum conditions under transient load conditions even when operators are not in attendance.

Lubricant Management Most lubricants are immiscible in ammonia and separate out of the liquid easily when flow velocity is low or when temperatures are lowered. Normally, lubricants can be easily drained from the system. However, if the temperature is very low and the lubricant is not properly selected, it becomes a gummy mass that prevents refrigerant controls from functioning, blocks flow passages, and fouls heat transfer surfaces. Proper lubricant selection and management is often the key to a properly functioning system. In two-stage systems, proper design usually calls for lubricant separators on both the high- and low-stage compressors. A properly designed coalescing separator can remove almost all the lubricant that is in droplet or aerosol form. Lubricant that reaches its saturation vapor pressure and becomes a vapor cannot be removed by a separator. Separators that can cool the discharge gas condense much of the vapor for consequent separation. Using lubricants that have very low vapor pressures below 80°C can minimize carryover to 2 or 3 mg/kg. Take care, however, to ensure that refrigerant is not condensed and fed back into the compressor or separator, where it can lower lubricity and cause compressor damage. In general, direct-expansion and liquid overfeed system evaporators have fewer lubricant return problems than do flooded system evaporators because refrigerant flows continuously at good velocities to sweep lubricant from the evaporator. Low-temperature systems using hot-gas defrost can also be designed to sweep lubricant out of the circuit each time the system defrosts. This reduces the possibility of coating the evaporator surface and hindering heat transfer. Flooded evaporators can promote lubricant build-up in the evaporator charge because they may only return refrigerant vapor back to the system. In ammonia systems, the lubricant is simply drained from the surge drum. At low temperatures, this procedure is difficult if the lubricant selected has a pour point above the evaporator temperature. Lubricant Removal from Ammonia Systems. Most lubricants are miscible with liquid ammonia only in very small proportions. The proportion decreases with the temperature, causing lubricant to separate. Ammonia evaporation increases the lubricant ratio, causing more lubricant to separate. Increased density causes the lubricant (saturated with ammonia at the existing pressure) to form a separate layer below the ammonia liquid. Unless lubricant is removed periodically or continuously from the point where it collects, it can cover the heat transfer surface in the evaporator, reducing performance. If gage lines or branches to level controls are taken from low points (or lubricant is allowed to accumulate), these lines will contain lubricant. The higher lubricant density is at a lower level than the ammonia liquid. Draining lubricant from a properly located collection point is not difficult unless the temperature is so low that the lubricant does not flow readily. In this case, keeping the receiver at a higher temperature may be beneficial. Alternatively, a lubricant with a lower pour point can be selected. Lubricant in the system is saturated with ammonia at the existing pressure. When the pressure is reduced, ammonia vapor separates, causing foaming. Draining lubricant from ammonia systems requires special care. Ammonia in lubricant foam normally starts to evaporate and produces a smell. Operators should be made aware of this. On systems where lubricant is drained from a still, a spring-loaded drain valve, which closes if the valve handle is released, should be installed.

Valves Fig. 31 Hot-Gas Injection Evaporator for Operations at Low Load

Stop Valves. These valves, also commonly called shutoff or isolation valves, are generally manually operated, although

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2014 ASHRAE Handbook—Refrigeration (SI)

motor-actuated units are available. ASHRAE Standard 15 requires these valves in the inlet and outlet lines to all condensers, compressors, and liquid receivers. Additional valves are installed on vessels, evaporators, and long lengths of pipe so they can be isolated in case of leaks and to facilitate pumping out for servicing and evacuation. Sections of liquid piping that can experience hydraulic lockup in normal operation must be protected with a relief device (preferably vented back into the system). Only qualified personnel should be allowed to operate stop valves. Installing globe-type stop valves with the valve stems horizontal lessens the chance (1) for dirt or scale to lodge on the valve seat or disk and cause it to leak or (2) for liquid or lubricant to pocket in the area below the seat. Wet suction return lines (recirculation system) should use angle valves or globe valves (with their stems horizontal) to reduce the possibility of liquid pockets and reduce pressure drop. Welded flanged or weld-in-line valves are desirable for all line sizes; however, screwed valves may be used for 32 mm and smaller lines. Ammonia globe and angle valves should have the following features: • • • • •

Soft seating surfaces for positive shutoff (no copper or copper alloy) Back seating to permit repacking the valve stem while in service Arrangement that allows packing to be tightened easily All-steel construction (preferable) Bolted bonnets above 25 mm, threaded bonnets for 25 mm and smaller

Consider seal cap valves in refrigerated areas and for all ammonia piping. To keep pressure drop to a minimum, consider angle valves (as opposed to globe valves). Control Valves. Pressure regulators, solenoid valves, check valves, gas-powered suction stop valves, and thermostatic expansion valves can be flanged for easy assembly and removal. Alternative weld-in line valves with nonwearing body parts are available. Valves 40 mm and larger should have socket- or butt-welded companion flanges. Smaller valves can have threaded companion flanges. A strainer should be used in front of self-contained control valves to protect them from pipe construction material and dirt. Solenoid Valves. Solenoid valve stems should be upright, with their coils protected from moisture. They should have flexible conduit connections, where allowed by codes, and an electric pilot light wired in parallel to indicate when the coil is energized. Solenoid valves for high-pressure liquid feed to evaporators should have soft seats for positive shutoff. Solenoid valves for other applications, such as in suction, hot-gas, or gravity feed lines, should be selected for the pressure and temperature of the fluid flowing and for the pressure drop available. Relief Valves. Safety valves must be provided in conformance with ASHRAE Standard 15 and Section VIII, Division 1, of the ASME Boiler and Pressure Vessel Code. For ammonia systems, IIAR Bulletin 109 also addresses the subject of safety valves. Dual relief valve arrangements allow testing of the relief valves (Figure 32). The three-way stop valve is constructed so that it is always open to one of the relief valves if the other is removed to be checked or repaired.

Isolated Line Sections Sections of piping that can be isolated between hand valves or check valves can be subjected to extreme hydraulic pressures if cold liquid refrigerant is trapped in them and subsequently warmed. Additional pressure-relieving valves for such piping must be provided.

Insulation and Vapor Retarders Chapter 10 covers insulation and vapor retarders. Insulation and effective vapor retarders on low-temperature systems are very important. At low temperatures, the smallest leak in the vapor retarder can allow ice to form inside the insulation, which can totally destroy

Fig. 32 Dual Relief Valve Fitting for Ammonia

Fig. 33 Shell-and-Coil Economizer Arrangement the integrity of the entire insulation system. The result can significantly increase load and power usage.

SYSTEMS In selecting an engineered ammonia refrigeration system, several design options must be considered, including compression type (single stage, economized, or multistage), direct-expansion feed type (direct expansion, flooded, or liquid recirculation), and secondary coolants selection.

SINGLE-STAGE SYSTEMS The basic single-stage system consists of evaporator(s), a compressor, a condenser, a refrigerant receiver (if used), and a refrigerant control device (expansion valve, float, etc.). Chapter 2 of the 2013 ASHRAE Handbook—Fundamentals discusses the compression refrigeration cycle.

ECONOMIZED SYSTEMS Economized systems are frequently used with rotary screw compressors. Figure 33 shows an arrangement of the basic components, and Figure 34 shows an economizer/receiver with a screw compressor. Subcooling the liquid refrigerant before it reaches the evaporator reduces its enthalpy, resulting in a higher net refrigerating effect. Economizing is beneficial because the vapor generated during

Ammonia Refrigeration Systems

Fig. 34 Screw Compressor with Economizer/Receiver

Fig. 35 Two-Stage System with High- and Low-Temperature Loads subcooling is injected into the compressor partway through its compression cycle and must be compressed only from the economizer port pressure (which is higher than suction pressure) to the discharge pressure. This produces additional refrigerating capacity with less increase in unit energy input. Economizing is most beneficial at high pressure ratios. Under most conditions, economizing can provide operating efficiencies that approach that of two-stage systems, but with much less complexity and simpler maintenance. Economized systems for variable loads should be selected carefully. At approximately 75% capacity, most screw compressors revert to single-stage performance as the slide valve moves and opens the economizer port to the compressor suction area. A flash economizer, which is somewhat more efficient, may be used instead of the shell-and-coil economizer (Figure 33). However, ammonia liquid delivery pressure is reduced to economizer pressure. Additionally, the liquid is saturated at the lower pressure and subject to flashing with any pressure drop unless another means of subcooling is incorporated.

MULTISTAGE SYSTEMS Multistage systems compress gas from the evaporator to the condenser in several stages. They are used to produce temperatures of –25°C and below. This is not economical with single-stage compression. Single-stage reciprocating compression systems are generally limited to between 35 and 70 kPa (gage) suction pressure. With lubricant-injected economized rotary screw compressors, which have lower discharge temperatures because of the lubricant cooling, the low-suction temperature limit is about –40°C, but efficiency is very low. Two-stage systems are used down to about –60°C evaporator temperatures. Below this temperature, three-stage systems should be considered.

2.19 Two-stage systems consist of one or more compressors that operate at low suction pressure and discharge at intermediate pressure and have one or more compressors that operate at intermediate pressure and discharge to the condenser (Figure 35). Where either single- or two-stage compression systems can be used, two-stage systems require less power and have lower operating costs, but they can have a higher initial equipment cost. As pressure ratios increase, single-stage ammonia systems encounter problems such as (1) high discharge temperatures on reciprocating compressors causing lubricant to deteriorate, (2) loss of volumetric efficiency as high pressure leaks back to the lowpressure side through compressor clearances, and (3) excessive stresses on compressor moving parts. Thus, manufacturers usually limit the maximum pressure ratios for multicylinder reciprocating machines to approximately 7 to 9. For screw compressors, which incorporate cooling, compression ratio is not a limitation, but efficiency deteriorates at high ratios. When the overall system pressure ratio (absolute discharge pressure divided by absolute suction pressure) begins to exceed these limits, the pressure ratio across the compressor must be reduced. This is usually done by using a multistage system. A properly designed two-stage system exposes each of the two compressors to a pressure ratio approximately equal to the square root of the overall pressure ratio. In a three-stage system, each compressor is exposed to a pressure ratio approximately equal to the cube root of the overall ratio. When screw compressors are used, this calculation does not always guarantee the most efficient system. Another advantage to multistaging is that successively subcooling liquid at each stage of compression increases overall system operating efficiency. Additionally, multistaging can accommodate multiple loads at different suction pressures and temperatures in the same refrigeration system. In some cases, two stages of compression can be contained in a single compressor, such as an internally compounded reciprocating compressor. In these units, one or more cylinders are isolated from the others so they can act as independent stages of compression. Internally compounded compressors are economical for small systems that require low temperature.

Two-Stage Screw Compressor System A typical two-stage, two-temperature screw compressor system provides refrigeration for high- and low-temperature loads (Figure 36). For example, the high-temperature stage supplies refrigerant to all process areas operating between –2 and 10°C. A –8°C intermediate suction temperature is selected. The low-temperature stage requires a –37°C suction temperature for blast freezers and continuous or spiral freezers. The system uses a flash intercooler that doubles as a recirculator for the –8°C load. It is the most efficient system available if the screw compressor uses indirect lubricant cooling. If refrigerant injection cooling is used, system efficiency decreases. This system is efficient for several reasons: • Approximately 50% of the booster (low-stage) motor heat is removed from the high-stage compressor load by the thermosiphon lubricant cooler. Note: In any system, thermosiphon lubricant cooling for booster and high-stage compressors is about 10% more efficient than injection cooling. Also, plants with a piggyback, two-stage screw compressor system without intercooling or injection cooling can be converted to a multistage system with indirect cooling to increase system efficiency approximately 15%. • Flash intercoolers are more efficient than shell-and-coil intercoolers by several percent. • Thermosiphon lubricant cooling of the high-stage screw compressor provides the highest efficiency available. Installing indirect

2.20

2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 36 Compound Ammonia System with Screw Compressor Thermosiphon Cooled cooling in plants with liquid injection cooling of screw compressors can increase compressor efficiency by 3 to 4%. • Thermosiphon cooling saves 20 to 30% in electric energy during low-temperature months. When outside air temperature is low, the condensing pressure can be decreased to 600 to 700 kPa (gage) in most ammonia systems. With liquid injection cooling, the condensing pressure can only be reduced to approximately 850 to 900 kPa (gage). • Variable-Vi compressors with microprocessor control require less total energy when used as high-stage compressors. The controller tracks compressor operating conditions to take advantage of ambient conditions as well as variations in load.

Converting Single-Stage into Two-Stage Systems When plant refrigeration capacity must be increased and the system is operating below about 70 kPa (gage) suction pressure, it is usually more economical to increase capacity by adding a compressor to operate as the low-stage compressor of a two-stage system than to implement a general capacity increase. The existing single-stage compressor then becomes the high-stage compressor of the two-stage system. When converting, consider the following: • The motor on the existing single-stage compressor may have to be increased in size when used at a higher suction pressure. • The suction trap should be checked for sizing at the increased gas flow rate. • An intercooler should be added to cool the low-stage compressor discharge gas and to cool high-pressure liquid. • A condenser may need to be added to handle the increased condensing load. • A means of purging air should be added if plant suction gage pressure is below zero. • A means of automatically reducing compressor capacity should be added so that the system will operate satisfactorily at reduced system capacity points.

LIQUID RECIRCULATION SYSTEMS The following discussion gives an overview of liquid recirculation (liquid overfeed) systems. See Chapter 4 for more complete

information. For additional engineering details on liquid overfeed systems, refer to Stoecker (1988). In a liquid ammonia recirculation system, a pump circulates ammonia from a low-pressure receiver to the evaporators. The low-pressure receiver is a shell for storing refrigerant at low pressure and is used to supply evaporators with refrigerant, either by gravity or by a mechanical pump. It also takes suction from the evaporators and separates gas from the liquid. Because the amount of liquid fed into the evaporator is usually several times the amount that actually evaporates there, liquid is always present in the suction return to the low-pressure receiver. Frequently, three times the evaporated amount is circulated through the evaporator (see Chapter 4). Generally, the liquid ammonia pump is sized by the flow rate required and a pressure differential of about 170 kPa. This is satisfactory for most single-story installations. If there is a static lift on the pump discharge, the differential is increased accordingly. Additional pressure differential consideration should be given when evaporator pressures are maintained higher than the low-pressure receiver’s operating pressure. The low-pressure receiver should be sized by the cross-sectional area required to separate liquid and gas and by the volume between the normal and alarm liquid levels in the low-pressure receiver. This volume should be sufficient to contain the maximum fluctuation in liquid from the various load conditions (see Chapter 4). Liquid at the pump discharge is in the subcooled region. A total pressure drop of about 35 kPa in the piping can be tolerated. The remaining pressure is expended through the control valve and coil. Pressure drop and heat pickup in the liquid supply line should be low enough to prevent flashing in the liquid supply line. Provisions for liquid relief in the liquid main downstream of the pump check valve back to the low-pressure receiver are required, so when liquid-line solenoid valves at the various evaporators are closed, either for defrosting or for temperature control, the excess liquid can be relieved back to the receiver. Additionally, liquid relief is required ahead of the pump discharge check valve. Generally, relief regulators used for this purpose are set at about 275 kPa differential when positive-displacement pumps are used. When

Ammonia Refrigeration Systems centrifugal pumps are used, a hand expansion valve or a minimum flow orifice is acceptable to ensure that the pump is not dead-headed. The suction header between evaporators and low-pressure receiver should be pitched down at least 1% to allow excess liquid flow back to the low-pressure receiver. The header should be designed to avoid traps. Liquid Recirculation in Single-Stage System. Figure 37 shows the piping of a typical single-stage system with a low-pressure receiver and liquid ammonia recirculation feed.

Hot-Gas Defrost This section was taken from a technical paper by Briley and Lyons (1992). Several methods are used for defrosting coils in areas below 2°C room temperature: • Hot refrigerant gas (the predominant method) • Water • Air • Combinations of hot gas, water, and air The evaporator (air unit) in a liquid recirculation system is circuited so that the refrigerant flow provides maximum cooling efficiency. The evaporator can also work as a condenser if the necessary piping and flow modifications are made. When the evaporator operates as a condenser and the fans are shut down, hot refrigerant vapor raises the surface temperature of the coil enough to melt any ice and/ or frost on the surface so that it drains off. Although this method is effective, it can be troublesome and inefficient if the piping system is not properly designed. Even when fans are not operating, 50% or more of the heat given up by the refrigerant vapor may be lost to the space. Because the heat transfer rate varies with the temperature difference between coil surface and room air, the temperature/pressure of the refrigerant during defrost should be minimized. Another reason to maintain the lowest possible defrost temperature/pressure, particularly in freezers, is to keep the coil from steaming. Steam increases refrigeration load, and the resulting icicle or frost formation must be dealt with. Icicles increase maintenance during cleanup; ice formed during defrost tends to collect at the fan rings, which sometimes restricts fan operation. Defrosting takes slightly longer at lower defrost pressures. The shorter the time heat is added to the space, the more efficient the defrost. However, with slightly extended defrost times at lower temperature, overall defrosting efficiency is much greater than at

2.21 higher temperature/pressure because refrigeration requirements are reduced. Another loss during defrost can occur when hot or uncondensed gas blows through the coil and relief regulator and vents back to the compressor. Some of this gas load cannot be contained and must be vented to the compressor through the wet return line. It is most energy efficient to vent this hot gas to the highest suction possible; an evaporator defrost relief should be vented to the intermediate or high-stage compressor if the system is two-stage. Figure 38 shows a conventional hot-gas defrost system for evaporator coils of 50 kW of refrigeration and below. Note that the wet return is above the evaporator and that a single riser is used. Defrost Control. Because defrosting efficiency is low, frequency and duration of defrosting should be kept to the minimum necessary to keep the coils clean. Less defrosting is required during winter than during hotter, more humid periods. An effective energysaving measure is to reset defrost schedules in the winter. Several methods are used to initiate the defrost cycle. Demand defrost, actuated by a pressure device that measures air pressure drop across the coil, is a good way of minimizing total daily defrost time. The coil is defrosted automatically only when necessary. Demand initiation, together with a float drainer to dump the liquid formed during defrost to an intermediate vessel, is the most efficient defrost system available (Figure 39). The most common defrost control method, however, is timeinitiated, time-terminated; it includes adjustable defrost duration and an adjustable number of defrost cycles per 24 h period. This control is commonly provided by a defrost timer. Estimates indicate that the load placed on a refrigeration system by a coil during defrost is up to three times the operating design load. Although estimates indicate that the maximum hot-gas flow can be up to three times the normal refrigeration flow, note that the hot-gas flow varies during the defrost period because of the amount of ice remaining on the coils. Hot-gas flow is greatest at the beginning of the defrost period, and decreases as the ice melts and the coil warms. It is therefore not necessary to engineer for the maximum flow, but for some lesser amount. The lower flow imposed by reducing the hot-gas pipe and valve sizes reduces the maximum hot-gas flow rate and makes the system less vulnerable to various shocks. Estimates show that engineering for hot-gas flow rates equal to the normal refrigeration flow rate is adequate and only adds a small amount of time to the overall defrost period to achieve a clean defrost.

Fig. 37 Piping for Single-Stage System with Low-Pressure Receiver and Liquid Ammonia Recirculation

2.22

2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 38

Conventional Hot-Gas Defrost Cycle

(For coils with 50 kW refrigeration capacity and below)

Fig. 39 Demand Defrost Cycle (For coils with 50 kW refrigeration capacity and below)

Ammonia Refrigeration Systems Designing Hot-Gas Defrost Systems. Several approaches are followed in designing hot-gas defrost systems. Figure 39 shows a typical demand defrost system for both upfeed and downfeed coils. This design returns defrost liquid to the system’s intermediate pressure. An alternative is to direct defrost liquid into the wet suction. A float drainer or thermostatic trap with a hot-gas regulator installed at the hot-gas inlet to the coil is an alternative to the relief regulator (see Figure 39). When using a condensate drainer, the device must never be allowed to stop the flow completely during defrost, because this allows the condensed hot gas remaining in the coil to pool in the lower circuits and become cold. Once this happens, defrosting of the lower circuits ceases. Water still running off the upper circuits refreezes on the lower circuits, resulting in ice build-up over successive defrosts. Any condensate drainer that can cycle closed when condensate flow momentarily stops should be bypassed with a metering valve or an orifice. Most defrost systems installed today (Figure 38) use a time clock to initiate defrost; the demand defrost system shown in Figure 39 uses a low-differential-pressure switch to sense the air pressure drop across the coil and actuate the defrost. A thermostat terminates the defrost cycle. A timer is used as a back-up to ensure the defrost terminates. Sizing and Designing Hot-Gas Piping. Hot gas is supplied to the evaporators in two ways: • The preferred method is to install a pressure regulator set at approximately 700 kPa (gage) in the equipment room at the hotgas takeoff and size the piping accordingly. • The alternative is to install a pressure regulator at each evaporator or group of evaporators and size the piping for minimum design condensing pressure, which should be set such that the pressure at the outlet of the coil is approximately 480 kPa (gage). This normally requires the regulator installed at the coil inlet to be set to about 620 kPa (gage). A maximum of one-third of the coils in a system should be defrosted at one time. If a system has 900 kW of refrigeration capacity, the main hot-gas supply pipe could be sized for 300 kW of refrigeration. Hot-gas mains should be sized one pipe size larger than the values given in Table 3 for hot-gas branch lines under 30 m. The outlet pressure-regulating valve should be sized in accordance with the manufacturer’s data. Reducing defrost hot-gas pressure in the equipment room has advantages, notably that less liquid condenses in the hot-gas line as the condensing temperature drops to 11 to 18°C. A typical equipment room hot-gas pressure control system is shown in Figure 40. If hot-gas lines in the system are trapped, a condensate drainer must be installed at each trap and at the low point in the hot-gas line (Figure 41). Defrost condensate liquid return piping from coils where a float or thermostatic valve is used should be one size larger than the liquid feed piping to the coil. Hot-gas defrost systems can be subject to hydraulic shock. See the section on Avoiding Hydraulic Shock, under Safety Considerations. Demand Defrost. The following are advantages and features of demand defrost: • It uses the least energy for defrost. • It increases total system efficiency because coils are off-line for a minimum amount of time. • It imposes less stress on the piping system because there are fewer defrost cycles. Soft Hot-Gas Defrost System. This system is particularly well suited to large evaporators and should be used on all coils of 50 kW of refrigeration or over. It eliminates the valve clatter, pipe movements, and some of the noise associated with large coils during hotgas defrost. Soft hot-gas defrost can be used for upfeed or downfeed

2.23

Fig. 40

Equipment Room Hot-Gas Pressure Control System

Fig. 41

Hot-Gas Condensate Return Drainer

coils; however, the piping systems differ (Figure 42). Coils operated in the horizontal plane with vertical headers must be orificed. Vertical coils with horizontal headers that usually are crossfed are also orificed. Soft hot-gas defrost is designed to increase coil pressure gradually as defrost begins. This is accomplished by a small hot-gas feed having a capacity of about 25 to 30% of the estimated duty with a solenoid and a hand expansion valve adjusted to bring the pressure up to about 275 kPa (gage) in 3 to 5 min. (See Sequence of Operation in Figure 42.) After defrost, a small suction-line solenoid is opened so that the coil can be brought down to operation pressure gradually before liquid is introduced and the fans started. The system can be initiated by a pressure switch; however, for large coils in spiral or individual quick freezing systems, manual initiation is preferred. Note that control valves are available to provide the soft-gas feature in combination with the main hot-gas valve capacity. There are also combination suction valves to provide pressure bleeddown at the end of the defrost cycle. The following additional features can make a soft hot-gas defrost system operate more smoothly and help avoid shocks to the system: • Regulating hot gas to approximately 725 kPa (gage) in the equipment room gives the gas less chance of condensing in supply piping. Liquid in hot-gas systems may cause problems because of the hydraulic shock created when the liquid is accelerated into an evaporator (coil). Coil headers and pan coils may rupture as a result. • Draining condensate formed during the defrost period with a float or thermostatic drainer eliminates hot-gas blowby normally associated with pressure-regulating valves installed around the wet suction return line pilot-operated check valve. • Returning liquid ammonia to the intercooler or high-stage recirculator saves considerable energy. A 70 kW refrigeration coil defrosting for 12 min can condense up to 11 kg/min of ammonia, or 132 kg total. The enthalpy difference between returning to the low-stage recirculator (–40°C) and the intermediate recirculator (–7°C) is 148 kJ/kg, for 19.5 MJ total or 27 kW of refrigeration removed from the –40°C booster for 12 min. This assumes that only liquid is drained and is the saving when liquid is drained to the

2.24

2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 42 Soft Hot-Gas Defrost Cycle (For coils with 50 kW refrigeration capacity or above)

intermediate point, not the total cost to defrost. If a pressure-regulating valve is used around the pilot-operated check valve, this rate could double or triple because hot gas flows through these valves in greater quantities. Soft hot-gas defrost systems reduce the probability of experiencing hydraulic shock. See the section on Avoiding Hydraulic Shock, under Safety Considerations. This system eliminates check valve chatter and most, if not all, liquid hammer (i.e., hydraulic problems in the piping). In addition, the last three features listed in the section on Demand Defrost apply to soft hot-gas defrost.

Double-Riser Designs for Large Evaporator Coils Static pressure penalty is the pressure/temperature loss associated with a refrigerant vapor stream bubbling through a liquid bath. If speed in the riser is high enough, it will carry over a certain amount of liquid, thus reducing the penalty. For example, at –40°C ammonia has a density of 689.9 kg/m3, which is equivalent to a pressure of 689.9(9.807 m/s2)/1000 = 6.77 kPa per metre of depth. Thus, a 5 m riser has a column of liquid that exerts 5  6.77 = 33.9 kPa. At –40°C, ammonia has a saturation pressure of 71.7 kPa. At the bottom of the riser then, the pressure is 33.9 + 71.7 = 105.6 kPa, which is the saturation pressure of ammonia at –33°C. This 7 K difference amounts to a 1.4 K penalty per metre of riser. If a riser were oversized to the point that the vapor did not carry liquid to the wet return, the evaporator would be at –33°C instead of –40°C. This problem can be solved in several ways:

Fig. 43 Recirculated Liquid Return System • Install the low-temperature recirculated suction (LTRS) line below the evaporator. This method is very effective for downfeed evaporators. Suction from the coil should not be trapped. This arrangement also ensures lubricant return to the recirculator. • Where the LTRS is above the evaporator, install a liquid return system below the evaporator (Figure 43). This arrangement eliminates static penalty, which is particularly advantageous for plate, individual quick freeze, and spiral freezers. • Use double risers from the evaporator to the LTRS (Figure 44). If a single riser is sized for minimum pressure drop at full load, the static pressure penalty is excessive at part load, and lubricant return

Ammonia Refrigeration Systems

2.25 Avoiding Hydraulic Shock

Fig. 44

Double Low-Temperature Suction Risers

could be a problem. If the single riser is sized for minimum load, then riser pressure drop is excessive and counterproductive. Double risers solve these problems (Miller 1979). Figure 44 shows that, when maximum load occurs, both risers return vapor and liquid to the wet suction. At minimum load, the large riser is sealed by liquid ammonia in the large trap, and refrigerant vapor flows through the small riser. A small trap on the small riser ensures that some lubricant and liquid return to the wet suction. Risers should be sized so that pressure drop, calculated on a dry-gas basis, is at least 70 Pa/m. The larger riser is designed for approximately 65 to 75% of the flow and the small one for the remainder. This design results in a velocity of approximately 25 m/ s or higher. Some coils may require three risers (large, medium, and small). Over the years, freezer capacity has grown. As freezers became larger, so did the evaporators (coils). Where these freezers are in line and the product to be frozen is wet, the defrost cycle can be every 4 or 8 h. Many production lines limit defrost duration to 30 min. If coils are large (some coils have a refrigeration capacity of 700 to 1000 kW), it is difficult to design a hot-gas defrost system that can complete a safe defrost in 30 min. Sequential defrost systems, where coils are defrosted alternately during production, are feasible but require special treatment.

SAFETY CONSIDERATIONS Ammonia is an economical choice for industrial systems. Although ammonia has superior thermodynamic properties, it is considered toxic at low concentration levels of 35 to 50 mg/kg. Large quantities of ammonia should not be vented to enclosed areas near open flames or heavy sparks. Ammonia at 16 to 25% by volume burns and can explode in air in the presence of an open flame. The importance of ammonia piping is sometimes minimized when the main emphasis is on selecting major equipment pieces. Liquid and suction mains should be sized generously to provide low pressure drop and avoid capacity or power penalties caused by inadequate piping. Hot-gas mains, on the other hand, should be sized conservatively to control the peak flow rates. In a large system with many evaporators, not all of them defrost simultaneously, so mains should only be engineered to provide sufficient hot gas for the number and size of coils that will defrost concurrently. Slight undersizing of the hot-gas piping is generally not a concern because the period of peak flow is short and the defrost cycles of different coils can be staggered. The benefit of smaller hot-gas piping is that the mass of any slugs that form in the piping is smaller.

Cold liquid refrigerant should not be confined between closed valves in a pipe where the liquid can warm and expand to burst piping components. Hydraulic shock, also known as water hammer, occurs in twophase systems experiencing pressure changes. Most engineers are familiar with single-phase water hammer, as experienced in water systems or occasionally in the liquid lines of refrigeration systems. These shocks, though noisy, are not widely known to cause damage in refrigeration systems. Damaging hydraulic shock events are almost always of the condensation-induced type. They occur most frequently in low-temperature ammonia systems and are often associated with the onset or termination of hot-gas defrosting. Failed system components are frequently evaporators, hot-gas inlet piping components associated with the evaporators, or two-phase suction piping and headers exiting the evaporators. Although hydraulic shock piping failures occur suddenly, there are usually reports of previous noise at the location of the failed component associated with hot-gas defrosting. ASHRAE research project RP-970 (Martin et al. 2008) found that condensation-induced hydraulic shocks are the result of liquid slugs in two-phase sections of the piping or equipment. The slugs normally do not occur during the refrigeration cycle or the hot-gas defrost cycle, but during the transition from refrigeration to hot gas or back. During the transitions, pressure in the evaporator rises at the beginning of the cycle (i.e., gas from the system’s high side rushes into the low side), and is relieved at the end (i.e., gas rushes out into the suction side). At the beginning of these transitions, the pressure imbalances are at their maximums, generating the highest gas flows. If the gas flows are sufficiently large, they scoop up liquid from traps or the bottom of two-phase pipes. Once the slug forms, it begins to compress the gas in front of it. If this gas is pushed into a partially filled evaporator or a section of piping without an exit (e.g., the end of a suction header), it will compress even more. Compression raises the saturation temperature of the gas to a point where it starts to condense on the cold piping and cold liquid ammonia. Martin et al. (2008) found that this condensation maintained a reasonably fixed pressure difference across the slug, and that the slug maintained a reasonably constant speed along the 6 m of straight test pipe. In tests where slugs occurred, pressure differentials across the slugs varied from about 35 to 70 kPa, and slug speeds from about 6 to 17 m/s. These slugs caused hydraulic shock peak pressures of as much as 5.2 MPa (gage). Conditions that are most conducive to development of hydraulic shock in ammonia systems are suction pressures below 35 kPa (gage) and defrost pressures of 480 kPa (gage) or more. During the transition from refrigeration to defrost, liquid slugs can form in the hot-gas piping. If the evaporator or its inlet hot-gas piping are not thoroughly drained before defrosting begins, the slugs will impact the standing liquid in the undrained evaporator and cause shocks, possibly damaging the evaporator or its hot-gas inlet piping. During the transition from defrost back to refrigeration, the 480+ kPa (gage) gas in the evaporator is released into the suction piping. Liquid slugs can come from traps in the suction piping or by picking up slower-moving liquid in wet suction piping. These slugs can be dissipated at suction line surge vessels, but if the suction piping arrangement is such that an inlet to a dead-end section of piping becomes sealed, and the dead-end section is sufficiently long compared to its diameter, then a shock can occur as gas in the dead-end section condenses and draws liquid into the section behind it. The shock occurs when the gas is all condensed and the liquid hits the closure (e.g., an end cap or a valve in the off position). This type of shock has been known to occur in piping as large as 400 mm. Low-temperature double pumper drum and low-temperature gaspowered transfer systems can also be prone to hydraulic shocks,

2.26 because these systems use hot gas to move low-temperature liquid. If slugs form in the gas lines or gas is pumped into the liquid lines, then there is potential for hydraulic shock: trapped gas can condense, causing the liquid to impact a closed valve or other piping element. To decrease the possibility of hydraulic shocks in ammonia systems, adhere to the following engineering guidelines: • Hot-gas piping should include no liquid traps. If traps are unavoidable, they should be equipped with liquid drainers. • If hot-gas piping is installed in cold areas of the plant or outdoors, the hot-gas condensate that forms in the piping should be drained and prevented from affecting the evaporator when the hot-gas valve opens. • The evaporator must be fully drained before opening the hot-gas valve, giving any liquid slugs in the hot gas free flow through the evaporator to the suction piping. If the liquid slugs encounter standing liquid in the evaporator, such as in the vertical evaporator suction header of an upfeed coil, shocks can occur. • Close attention should be paid to initial and sustained hot-gas flow rates when sizing control valves and designing the control valve assemblies. Emphasize keeping hot-gas piping and valves as small as possible, to reduce the peak mass flow rate of the hot gas. • Evaporator shutoff valves should be installed with their stems horizontal. • Wet suction lines should contain no traps, except for the trap in a double riser assembly. Between each evaporator and the low-pressure receiver, there should be no more than one high point in the piping. This means that the suction branch to each evaporator should contain a high point located above the suction main. • Wet suction mains and branches should contain no dead-end sections. Be especially careful with valved crossovers between parallel suction lines, because these become dead ends when the valve is closed. • In liquid transfer vessels or the vessels of double pumper systems, take extra precautions to ensure that the liquid level is maintained between the 20% and 80% full marks. Draining a vessel or overfilling puts gas in liquid lines or liquid in gas lines, and can cause hydraulic shock.

Hazards Related to System Cleanliness Rusting pipes and vessels in older systems containing ammonia can create a safety hazard. Oblique x-ray photographs of welded pipe joints and ultrasonic inspection of vessels may be used to disclose defects. Only vendor-certified parts for pipe, valving, and pressure-containing components according to designated assembly drawings should be used to reduce hazards. Most service problems are caused by inadequate precautions during design, construction, and installation (ASHRAE Standard 15; IIAR Standard 2). Ammonia is a powerful solvent that removes dirt, scale, sand, or moisture remaining in the pipes, valves, and fittings during installation. These substances are swept along with the suction gas to the compressor, where they are a menace to the bearings, pistons, cylinder walls, valves, and lubricant. Most compressors are equipped with suction strainers and/or additional disposable strainer liners for the large quantity of debris that can be present at initial start-up. Moving parts are often scored when a compressor is run for the first time. Damage starts with minor scratches, which increase progressively until they seriously affect compressor operation or render it inoperative. A system that has been carefully and properly installed with no foreign matter or liquid entering the compressor will operate satisfactorily for a long time. As piping is installed, it should be power rotary wire brushed and blown out with compressed air. The piping system should be blown out again with compressed air or nitrogen

2014 ASHRAE Handbook—Refrigeration (SI) before evacuation and charging. See ASHRAE Standard 15 for system piping test pressure.

REFERENCES ASHRAE. 2002. ASHRAE position document: Ammonia as a refrigerant. ASHRAE. 2007. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2007. ASME. 2010. Rules for construction of pressure vessels. Boiler and pressure vessel code, Section VIII, Division 1. American Society of Mechanical Engineers, New York. ASME. 2010. Refrigeration piping and heat transfer components. ANSI/ ASME Standard B31.5-2010. American Society of Mechanical Engineers, New York. ASTM. 2007. Specification for pipe, steel, black and hot-dipped, zinccoated, welded and seamless. ANSI/ASTM Standard A53/A53M-07. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2008. Specification for seamless carbon steel pipe for high-temperature service. ANSI/ASTM Standard A106/A106M-08. American Society for Testing and Materials, West Conshohocken, PA. Balmer, R.T. 2010. Modern engineering thermodynamics, p. 548. Academic Press, Waltham, MA. Briley, G.C., and T.A. Lyons. 1992. Hot gas defrost systems for large evaporators in ammonia liquid overfeed systems. IIAR Technical Paper 163. International Institute of Ammonia Refrigeration, Arlington, VA. Dinçer, I. 1997. Heat transfer in food cooling applications, p. 125. Taylor & Francis, Washington, D.C. Frick Co. 2004. Thermosyphon oil cooling. Bulletin E70-90E (July). Frick Company, Waynesboro, PA. GPO. 1893. United States Congressional serial set, vol. 41, p. 655. Government Printing Office, Washington D.C. IIAR. No date. Ammonia: The natural refrigerant of choice. International Institute of Ammonia Refrigeration, Alexandria, VA. IIAR. 1992. Avoiding component failure in industrial refrigeration systems caused by abnormal pressure or shock. Bulletin 116. International Institute of Ammonia Refrigeration, Arlington, VA. IIAR. 1998. Minimum safety criteria for a safe ammonia refrigeration system. Bulletin 109. International Institute of Ammonia Refrigeration, Arlington, VA. IIAR. 2008. Equipment, design, and installation of ammonia mechanical refrigeration systems. ANSI/IIAR Standard 2-2008. International Institute of Ammonia Refrigeration, Arlington, VA. Loyko, L. 1992. Condensation induced hydraulic shock. IIAR Technical Paper. International Institute of Ammonia Refrigeration, Arlington, VA. Martin, C.S., R. Brown, J. Brown, L. Loyko, and R. Cole. 2008. Condensation-induced hydraulic shock laboratory study. ASHRAE Research Project RP-970, Final Report. Miller, D.K. 1979. Sizing dual-suction risers in liquid overfeed refrigeration systems. Chemical Engineering (September 24). NCPWB. No date. Welding procedure specifications. National Certified Pipe Welding Bureau, Rockville, MD. Schmidt, L.M. 1908. Principles and practice of artificial ice making and refrigeration, p. 194. Philadelphia Book Co. Shelton, J.C., and A.M. Jacobi. 1997a. A fundamental study of refrigerant line transients: Part 1—Description of the problem and survey of relevant literature. ASHRAE Transactions 103(1):65-87. Shelton, J.C., and A.M. Jacobi. 1997b. A fundamental study of refrigerant line transients: Part 2—Pressure excursion estimates and initiation mechanisms. ASHRAE Transactions 103(2):32-41. Stoecker, W.F. 1988. Industrial refrigeration, Chapters 8 and 9 in Business News, Troy, MI. Timm, M.L. 1991. An improved method for calculating refrigerant line pressure drops. ASHRAE Transactions 97(1):194-203. Wile, D.D. 1977. Refrigerant line sizing. Final Report, ASHRAE Research Project RP-185. Woolrich, W.R., and C.T. Clark. No date. Refrigeration. Texas State Historical Association.

BIBLIOGRAPHY ASHRAE. 2002. ASHRAE position document on ammonia as a refrigerant. Available at https://www.ashrae org/File Library/docLib/About Us /PositionDocuments/ASHRAE_PD _Ammonia_Refrigerant_2013.pdf.

Ammonia Refrigeration Systems BAC. 1983. Evaporative condenser engineering manual. Baltimore Aircoil Company, Baltimore, MD. Bradley, W.E. 1984. Piping evaporative condensers. In Proceedings of IIAR Meeting, Chicago. International Institute of Ammonia Refrigeration, Arlington, VA. Cole, R.A. 1986. Avoiding refrigeration condenser problems. Heating/ Piping/Air-Conditioning, Parts I and II, 58(7, 8). Dinçer, I. 1997. Heat transfer in food cooling applications. Taylor and Francis, Washington, D.C. Glennon, C., and R.A. Cole. 1998. Case study of hydraulic shock events in an ammonia refrigerating system. IIAR Technical Paper. International Institute of Ammonia Refrigeration, Arlington, VA.

2.27 Loyko, L. 1989. Hydraulic shock in ammonia systems. IIAR Technical Paper T-125. International Institute of Ammonia Refrigeration, Arlington, VA. Nickerson, J.F. 1915. The development of refrigeration in the United States. Ice and Refrigeration 49(4):170-177. Available at http://books.google .com/books?id=YZc7AQAAMAAJ. Nuckolls, A.H. The comparative life, fire, and explosion hazards of common refrigerants. Miscellaneous Hazard 2375. Underwriters Laboratory, Northbrook, IL. Strong, A.P. 1984. Hot gas defrost—A-one-a-more-a-time. IIAR Technical Paper T-53. International Institute of Ammonia Refrigeration, Arlington, VA. USGS. 2012. 2011 minerals yearbook. U.S. Geological Survey, Reston, VA.

CHAPTER 3

CARBON DIOXIDE REFRIGERATION SYSTEMS Applications ............................................................................... System Design ............................................................................ System Safety.............................................................................. Piping......................................................................................... Heat Exchangers and Vessels.....................................................

3.2 3.3 3.7 3.7 3.8

Compressors for CO2 Refrigeration Systems ............................. 3.8 Lubricants .................................................................................. 3.9 Evaporators................................................................................ 3.9 Defrost ...................................................................................... 3.10 Installation, Start-up, and Commissioning .............................. 3.11

ARBON dioxide (R-744) is one of the naturally occurring compounds collectively known as “natural refrigerants.” It is nonflammable and nontoxic, with no known carcinogenic, mutagenic, or other toxic effects, and no dangerous products of combustion. Using carbon dioxide in refrigerating systems can be considered a form of carbon capture, with a potential beneficial effect on climate change. It has no adverse local environmental effects. Carbon dioxide exists in a gaseous state at normal temperatures and pressures within the Earth’s atmosphere. Currently, the global average concentration of CO2 is approximately 390 ppm by volume. Carbon dioxide has a long history as a refrigerant. Since the 1860s, the properties of this natural refrigerant have been studied and tested in refrigeration systems. In the early days of mechanical refrigeration, few suitable chemical compounds were available as refrigerants, and equipment available for refrigeration use was limited. Widespread availability made CO2 an attractive refrigerant. The use of CO2 refrigeration systems became established in the 1890s and CO2 became the refrigerant of choice for freezing and transporting perishable food products around the world. Meat and other food products from Argentina, New Zealand and Australia were shipped via refrigerated vessels to Europe for distribution and consumption. Despite having traveled a several-week voyage spanning half the globe, the receiving consumer considered the condition of the frozen meat to be comparable to the fresh product. By 1900, over 300 refrigerated ships were delivering meat products from many distant shores. In the same year, Great Britain imported 360,000 tons of refrigerated beef and lamb from Argentina, New Zealand, and Australia. The following year, refrigerated banana ships arrived from Jamaica, and tropical fruit became a lucrative cargo for vessel owners. CO2 gained dominance as a refrigerant in marine applications ranging from coolers and freezers for crew provisions to systems designed to preserve an entire cargo of frozen products. Safety was the fundamental reason for CO2’s development and growth. Marine CO2-refrigerated shipping rapidly gained popularity for its reliability in the distribution of a wide variety of fresh food products to many countries around the world. The CO2 marine refrigeration industry saw phenomenal growth, and by 1910 some 1800 systems were in operation on ships transporting refrigerated food products. By 1935, food producers shipped millions of tons of food products including meats, dairy products, and fruits to Great Britain annually. North America also was served by CO2 marine refrigeration in both exporting and receiving food products. The popularity of CO2 refrigeration systems reduced once suitable synthetic refrigerants became available. The development of chlorodifluoromethane (R-22) in the 1940s started a move away from CO2, and by the early 1960s it had been almost entirely replaced in all marine and land-based systems. By 1950, the chlorofluorocarbons (CFCs) dominated the majority of land-based refrigeration systems. This included a wide variety of domestic and commercial CFC uses. The development of the

hermetic and semihermetic compressors accelerated the development of systems containing CFCs. For the next 35 years, a number of CFC refrigerants gained popularity, replacing practically all other refrigerants except ammonia, which maintained its dominant position in industrial refrigeration systems. In the 1970s, the atmospheric effects of CFC emissions were highlighted. This lead to a concerted effort from governments, scientists, and industrialists to limit these effects. Initially, this took the form of quotas on production, but soon moved to a total phaseout, first of CFCs and then of hydrochlorofluorocarbons (HCFCs). The ozone depleting potential (ODP) rating of CFCs and HCFCs prompted the development of hydrofluorocarbon (HFC) refrigerants. Subsequent environmental research shifted the focus from ozone depletion to climate change, producing a second rating known as the global warming potential (GWP). Table 1 presents GWPs for several common refrigerants. Table 2 compares performance of current refrigerants used in refrigeration systems. In recent years, CO2 has once again become a refrigerant of great interest. However, high-pressure CO2 systems (e.g., 3.4 MPa at a saturation temperature of –1°C, or 6.7 MPa at 26.7°C) present some challenges for containment and safety. Advances in materials science since the 1950s enable the design of cost-effective and efficient high-pressure carbon dioxide systems. The attraction of using CO2 in modern systems is based on its

C

The preparation of this chapter is assigned to TC 10.3, Refrigerant Piping, Controls and Accessories.

3.1

Table 1 Refrigerant Data Refrigerant Refrigerant Number Group

Chemical Formula

R-22 HCFC R-134a HFC R-410A HFC blend

CHClF2 CF3CH2F HFC-32 (50%) HFC-125 (50%) R-507A HFC blend HFC-125 (50%) HFC-143a (50%) R-717 Ammonia NH3 R-744 Carbon dioxide CO2

Temperature GWP at at Safety 100 101.3 kPa, °C Group Years –40.8 –26.1 –52.3

A1 A1 A1/A1

1700 1300 2000

–47.1

A1

3900

–33.3 –78.4

B2 A1

0 1

Source: Adapted from ANSI/ASHRAE Standard 34-2007.

Table 2 Comparative Refrigerant Performance per Kilowatt of Refrigeration EvaporaConNet RefrigSpecific Refrigtor denser erating Refrigerant Volume of erant Pressure, Pressure, Effect, Circulated, Suction Gas, Number MPa MPa kJ/kg kg/s m3/kg R-22 R-134a R-410A R-507A R-717 R-744

0.3 0.16 0.48 0.38 0.24 2.25

1.19 0.77 1.87 1.46 1.16 7.18

162.2 147.6 167.6 110.0 1100.9 133.0

1.7 × 10–3 1.9 × 10–3 1.7 × 10–3 2.6 × 10–3 0.26 × 10–3 1.1 × 10–3

2.7 × 10–3 4.2 × 10–3 1.9 × 10–3 1.8 × 10–3 17.6 × 10–3 0.58 × 10–3

Source: Adapted from Table 9 in Chapter 29 of the 2009 ASHRAE Handbook—Fundamentals. Conditions are –15°C and 30°C.

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 1 CO2 Expansion-Phase Changes (Adapted from Vestergaard and Robinson 2003)

attractive thermophysical properties: low viscosity, high thermal conductivity, and high vapor density. These result in good heat transfer in evaporators, condensers, and gas coolers, allowing selection of smaller equipment compared to CFCs and HFCs. Carbon dioxide is unique as a refrigerant because it is being considered for applications spanning the HVAC&R market, ranging from freezers to heat pumps, and from domestic units up to large-scale industrial plants. CO2 has been proposed for use as the primary refrigerant in mobile air conditioners, domestic appliances, supermarket display cases, and vending machines. CO2 heat pump water heaters are already commercially available in a many countries. In these applications, transcritical operation (i.e., rejection of heat above the critical point) is beneficial because it allows good temperature glide matching between the water and supercritical CO2, which benefits the coefficient of performance (COP). Large industrial systems use CO2 as the low-temperature-stage refrigerant in cascade systems, typically with ammonia or R-507A as high-temperature-stage refrigerants. Medium-sized commercial systems also use CO2 as the low-temperature-stage refrigerant in cascade system with HFCs or hydrocarbons as high-temperature-stage refrigerants. A distinguishing characteristic of CO2 is its phase change properties. CO2 is commercially marketed in solid form as well as in liquid and gas cylinders. In solid form it is commonly called dry ice, and is used in a variety of ways including as a cooling agent and as a novelty or stage prop. Solid CO2 sublimates to gas at –78.5°C at atmospheric pressure. The latent heat is 571 kJ/kg. Gaseous CO2 is sold as a propellant and is available in high-pressure cartridges in capacities from 4 g to 2.3 m3. Liquid CO2 is dispensed and stored in large pressurized vessels that are often fitted with an independent refrigeration system to control storage vessel pressure. Manufacturing facilities use it in both liquid and gas phase, depending on the process or application. Bigger quantities of CO2 (e.g., to replenish large storage tanks) can be transported by pressurized railway containers and specialized road transport tanker trucks. CO2 is considered a very-low-cost refrigerant at just a fraction of the price of other common refrigerants in use today. Comparing environmental concerns, safety issues, and cost differentials, CO2 has a positive future in mechanical refrigeration systems, serving as both a primary and secondary refrigerant. In considering CO2 as primary or secondary refrigerant, these matter-phase state conditions of solid, liquid, and vapor should be thoroughly understood. Of particular importance are the triple point and critical point, which are illustrated in Figures 1 and 2. The point of equilibrium where all three states coexist that is known as the triple point. The second important pressure and temperature point of recognition is the critical point where liquid

Fig. 2 CO2 Phase Diagram (Adapted from Vestergaard and Robinson 2003)

and vapor change state. CO2 critical temperature is 31°C; this is considered to be low compared to all commonly used refrigerants.

APPLICATIONS Transcritical CO2 Refrigeration In a transcritical refrigeration cycle, CO2 is the sole refrigerant. Typical operating pressures are much higher than traditional HFC and ammonia operating pressures. As the name suggests, the heat source and heat sink temperatures straddle the critical temperature. Development on modern transcritical systems started in the early 1990s with a focus on mobile air-conditioning systems. However, early marine systems clearly were capable of transcritical operation in warm weather, according to their operating manuals. For example, marine engineers sailing through the Suez Canal in the 1920s reported that they had to throttle the “liquid” outlet from the condenser to achieve better efficiency if the sea water was too warm. They did not call this transcritical operation and could not explain why it was necessary, but their observation was correct. The technology suggested for mobile air conditioning was also adopted in the late 1990s for heat pumps, particularly air-source heat pumps for domestic water heating. In Japan, researchers and manufacturers have designed a full line of water-heating-system equipment, from small residential units to large industrial applications, all incorporating transcritical CO2 heat pump technology. A wide variety of such units was produced, with many different compressor types, including reciprocating, rotary piston, and scroll. Current commercial production of pure transcritical systems is primarily in small-scale or retail applications such as soft drink vending machines, mobile air conditioning, heat pumps, domestic appliances, and supermarket display freezers. Commercial and industrial systems at this time tend to use CO2 as secondary refrigerant in a two-phase cascade system in conjunction with more traditional primary refrigerants such as ammonia or an HFC. In a transcritical cycle, the compressor raises the operating pressure above the critical pressure and heat is rejected to atmosphere by cooling the discharge gas without condensation. When the cooled gas passes through an expansion device, it turns to a mixture of liquid and gas. If the compressor discharge pressure is raised, the enthalpy achieved at a given cold gas temperature is reduced, so there is an optimum operating point balancing the additional energy input required to deliver the higher discharge pressure against the additional cooling effect achieved through reduced enthalpy. Several optimizing algorithms have been developed to maximize efficiency by measuring saturated suction pressure and gas cooler outlet temperature and regulating the refrigerant flow to maintain an

Carbon Dioxide Refrigeration Systems optimum discharge pressure. Achieving as low a temperature at the gas cooler outlet as possible is key to good efficiency, suggesting that there is a need for evaporatively cooled gas coolers, although none are currently on the market. Other devices, such as expanders, have been developed to achieve the same effect by reducing the enthalpy during the expansion process and using the recovered work in the compressor to augment the electrical input.

CO2 Cascade System The cascade system consists of two independent refrigeration systems that share a common cascade heat exchanger. The CO2 lowtemperature refrigerant condenser serves as the high-temperature refrigerant evaporator; this thermally connects the two refrigeration circuits. System size influences the design of the cascade heat exchanger: large industrial refrigeration system may use a shelland-tube vessel, plate-and-frame heat exchanger, or plate-and-shell type, whereas commercial systems are more likely to use brazedplate, coaxial, and tube-in-tube cascade heat exchangers. In chilling systems, the liquid CO2 is pumped from the receiver vessel below the cascade heat exchanger to the heat load. In low-temperature applications, the high-pressure CO2 liquid is expanded to a lower pressure and a compressor is used to bring the suction gas back up to the condensing pressure. Using a cascade system allows a reduced high-temperature refrigerant charge. This can be important in industrial applications to minimize the amount of ammonia on site, or in commercial systems to reduce HFC refrigerant losses. CO2 cascade systems are configured for pumped liquid recirculation, direct expansion, volatile secondary and combinations of these that incorporate multiple liquid supply systems. Low-temperature cascade refrigeration application include cold storage facilities, plate freezers, ice machines, spiral and belt freezers, blast freezers, freeze drying, supermarkets, and many other food and industrial product freezing systems. Some theoretical studies [e.g., Vermeeren et al. (2006)] have suggested that cascade systems are inherently less efficient than twostage ammonia plants, but other system operators claim lower energy bills for their new CO2 systems compared to traditional ammonia plants. The theoretical studies are plausible because introducing an additional stage of heat transfer is bound to lower the high-stage compressor suction. However, additional factors such as the size of parasitic loads (e.g., oil pumps, hot gas leakage) on the low-stage compressors, the effect of suction line losses, and the adverse effect of oil in low-temperature ammonia plants all tend to offset the theoretical advantage of two-stage ammonia system, and in the aggregate the difference in energy consumption one way or the other is likely to be small. Other factors, such as reduced ammonia charge, simplified regulatory requirements, or reduced operator staff, are likely to be at least as significant in the decision whether to adopt CO2 cascades for industrial systems. In commercial installations, the greatest benefit of a CO2 cascade is the reduction in HFC inventory, and consequent probable reduction in HFC emission. Use of a cascade also enables the operator to retain existing HFC compressor and condenser equipment when refurbishing a facility by connecting it to a CO2 pump set and replacing the evaporators and low-side piping. End users in Europe and the United States suggest that CO2 cascade systems are simpler and easier to maintain, with fewer controls requiring adjustment, than the HFC systems that they are replacing. This indicates that they are inherently more reliable and probably cheaper to maintain than conventional systems. If the efficiency is equivalent, then the cost of ownership will ultimately be cheaper. However, it is not clear if these benefits derive from the higher level of engineering input required to introduce the new technology, or whether they can be maintained in the long term.

3.3 SYSTEM DESIGN Transcritical CO2 Systems Recent advances in system component design have made it possible to operate in previously unattainable pressure ranges. The development of hermetic and semihermetic multistage CO2 compressors provided the economical ability to design air-cooled transcritical systems that are efficient, reliable, and cost effective. Today, transcritical systems are commercially available in sizes from the smallest appliances to entire supermarket systems. Figures 3 and 4 shows examples of simple transcritical systems. Heat rejection to atmosphere is by cooling the supercritical CO2 gas without phase change. For maximum efficiency, the gas cooler must be able to operate as a condenser in colder weather, and the control system must be able to switch from gas cooler operation (where outflow from the air-cooled heat exchanger is restricted) to condenser operation (where the restriction is removed, as in a conventional system). Compared to a typical direct HFC system, energy usage can be reduced by 5% in colder climates such as northern Europe, but may

Fig. 3

Transcritical CO2 Refrigeration Cycle in Appliances and Vending Machines

Fig. 4 CO2 Heat Pump for Ambient Heat to Hot Water

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 5 R-717/CO2 Cascade System with CO2 Hot-Gas Defrosting (Adapted from Vestergaard 2007)

increase by 5% in warmer climates such as southern Europe or the United States. In a heat pump or a refrigeration system with heat recovery, this dual control is not necessary because the system operates transcritically at all times.

CO2/HFC Cascade Systems Cascade refrigeration systems in commercial applications generally use HFCs, or occasionally HCs, as the primary refrigerant. Supermarkets have adopted cascade technology for operational and economic reasons (the primary refrigerant charge can be reduced by as much as 75%). Liquid CO2 is pumped to low-temperature display cases and controlled via electronic expansion valve. The mediumtemperature display cases are supplied liquid from the same circuit or from a dedicated pump system (Figures 5 and 6). Cascade systems in supermarkets have been designed to operate multitemperature display cases and provide heat recovery to generate hot water or space heating (Figure 7). In general, although a pump has been introduced, energy consumption is not significantly different from a traditional HFC system because the suction line losses are less and the evaporator heat transfer performance is better. This can result in a rise of up to 5 or 6 K in the evaporating temperature, offsetting the pump’s power consumption and the temperature differential in the cascade heat exchanger.

Ammonia/CO2 Cascade Refrigeration System Industrial refrigeration applications often contain large amounts of ammonia as an operating charge. Cascade systems provide an opportunity to reduce the ammonia charge by approximately 90% percent compared to a conventional ammonia system of the same capacity.

Fig. 6 CO2 Cascade System with Two Temperature Levels (Adapted from Vestergaard 2007)

Another significant difference is the operating pressures of CO2 compared to ammonia. The typical suction pressure at –28.9°C evaporating temperature is 24.1 kPa (gage) for ammonia and 1582.4 kPa (gage) for CO2. In most industrial cascade systems, the ammonia charge is limited to the compressor room and the condenser flat, reducing the risk of leakage in production areas and cold storage rooms. The cascade heat exchanger is the main component where the two independent refrigeration systems are connected in single vessel. CO2 vapors are condensed to liquid by evaporating ammonia liquid to vapor. This cascade heat exchanger vessel must be constructed to withstand high pressures and temperature fluctuations to meet the requirements of both refrigerants. Also, the two refrigerants are not compatible with each other, and cross-contamination results in blockage in the ammonia circuit and may put the system out of commission for an extended period. The cascade heat exchanger design must prevent internal leakage that can lead to the two refrigerants reacting together. Figure 8 shows a simplified ammonia cascade system; note that no oil return is shown.

System Design Pressures The system design pressure for a CO2 cascade system cannot be determined in the traditional way, because the design temperatures are typically above the critical point. The system designer must therefore select suitable pressures for each part of the system, and ensure that the system is adequately protected against excess pressure in abnormal circumstances (e.g., off-cycle, downtime, power loss).

Carbon Dioxide Refrigeration Systems

3.5

Fig. 7 Dual-Temperature Supermarket System: R-404A and CO2 with Cascade Condenser For example, for a typical refrigerated warehouse or freezer cascade system, the following pressures are appropriate: CO2 Side • System design working pressure (saturated suction temperature): 3.5 MPa (gage) (0.6°C) • Relief valve settings: 3.4 MPa (gage) • System emergency relief setting: 3.1 MPa (gage) (–3°C) • CO2 discharge pressure setting: 2.2 MPa (gage) (–15°C) Where the system uses hot-gas defrost, the part of the circuit exposed to the high-pressure gas should be rated for 5.2 MPa or higher. Ammonia Side • System design working pressure (saturated suction temperature): 2.1 MPa (gage) (53°C) • Relief valve settings: 2.1 MPa (gage) • Ammonia suction pressure setting: 108 kPa (gage) (–18°C) • Ammonia discharge pressure setting: 1.1 MPa (gage) (32°C) • Temperature difference on the cascade condenser: 2.8 K On the CO2 side, the low-side temperature and coincident pressure must be considered. The triple point for CO2 is –56.6°C. At lower pressure, liquid turns to a solid; thus, the low-side criteria of feasible applications are –56.6°C at a coincidental saturated suction pressure of 414 kPa (gage). Therefore, the system must be dualstamped for 3.5 MPa (gage) and –56.6°C at 462 kPa (gage). To achieve suitable material properties, stainless steel pipe may be appropriate.

Valves Valves in CO2 systems are generally similar to those in ammonia plants, but must be suitably rated for high pressure. Where equipment cannot operate at the required pressure differences, alternative types may be used (e.g., replacing solenoid valves with electrically driven ball valves).

Expanding saturated CO2 vapor can solidify, depending on operating pressure, so the relief valve should be located outside with no downstream piping. If necessary, there should be a high-pressure pipe from the vessel to the relief valve. This pipe should be sized to ensure a suitably low pressure drop during full-flow operation. The other very important consideration with the relief system is its discharge location. The relief header must be located so that, if there is a release, the discharge does not fall and collect in an area where it may cause an asphyxiation hazard (e.g., in a courtyard, or near the inlet of a rooftop makeup air unit). CO2 relief valves are more likely to lift in abnormal circumstances than those used in ammonia or HFC systems, where the valve will only lift in the event of a fire or a hydraulic lock. Therefore, care should be taken when specifying relief valves for CO2 to ensure that the valve can reseat to prevent loss of the total refrigeration charge. A pressure-regulating valve (e.g., an actuated ball valve) may be installed in parallel with the safety relief valve to allow controlled venting of the vapor at a set pressure slightly lower than the relief valve setting. For sizing relief valves, use the following equation: C = f DL

(1)

where C D L f

= = = =

capacity required, kg/s of air diameter of vessel, m length of vessel, m refrigerant-specific constant (0.5 for ammonia, 1.0 for CO2)

Some special considerations are necessary for liquid feed valve assemblies to facilitate maintenance. Depending on the configuration, it may not be feasible to drain the liquid out of a valve assembly before maintenance is needed. Liquid CO2 in the valve assembly cannot be vented directly to atmosphere because it will turn to dry ice immediately. Between any two valves that can trap liquid, a liquid drain valve should be installed on one side and a gas-pressuring valve on the other. This facilitates pressurizing the valve train with gas, pushing the liquid out without it changing phase inside the pipe.

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2014 ASHRAE Handbook—Refrigeration (SI)

Fig. 8 Dual-Temperature Ammonia (R-717) Cascade System

CO2 Monitoring CO2 is heavier than air, but the two gases mix well; it does not take much air movement to prevent CO2 from stratifying. The most practical place to measure CO2 concentrations is about 1.2 m above the floor (i.e., the breathing zone for most people). Where CO2 might leak into a stairwell, pit, or other confined space, an additional detector should be located in the space to warn personnel in the event of a high concentration.

Water in CO2 Systems CO2, like HFCs, is very sensitive to any moisture within the system. Air must be evacuated before charging the refrigerant at initial start-up, to remove atmospheric moisture. Maintenance staff must use caution when adding oil that may contain moisture. Investigations of valve problems in some CO2 installations revealed that many problems are caused by water freezing in the system; welldesigned and well-maintained CO2 systems charged with dry CO2 and filter-driers run trouble free (Bellstedt et al. 2002). Figure 9 shows the water solubility in the vapor phase of different refrigerants. The acceptable level of water in CO2 systems is much lower than with other common refrigerants. Figure 10 shows the solubility of water in both liquid and vapor CO2 as function of

Fig. 9 Water Solubility in Various Refrigerants (Adapted from Vestergaard 2007)

temperature. (Note that solubility in the liquid phase is much higher.) Below these levels, water remains dissolved in the refrigerant and does not harm the system. If water is allowed to exceed the maximum solubility limit in a CO2 system, problems may occur, especially if the temperature is below 0°C. In this case, the water

Carbon Dioxide Refrigeration Systems

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Fig. 10 Water Solubility in CO2

ensuring that the relevant parts of the system have been relieved of pressure. When reducing pressure or transferring liquid carbon dioxide, care is necessary to guard against blockages caused by solid carbon dioxide, which forms at pressures below 517 kPa. If a blockage occurs, it must be treated with caution. No attempt should be made to accelerate the release of pressure by heating the blocked component. In a room where people are present and the CO2 concentration could exceed the refrigerant concentration limit of 0.9 kg/10 m3 in the event of a leak, proper detection and ventilation are required. When detectors sense a dangerous level of CO2 in a room, the alarm system must be designed to make sure all people in the room are evacuated and no one is allowed to reenter until concentration levels return to acceptable ranges. Protective clothing, including gloves and eyewear, should be standard in locations that contain CO2 equipment or controls, or where service work is done.

(Adapted from Vestergaard 2007)

PIPING freezes, and ice crystals may block control valves, solenoid valves, filters, and other equipment. If the water concentration in a CO2 system exceeds the saturation limit, it creates carbonic acid, which can cause equipment failures and possibly internal pipe corrosion. Filter-driers should be located at all main liquid feed locations. Because the entire CO2 system is at positive pressure during all operating conditions, the most likely time for moisture penetration is during charging. The appropriate specification for water content depends on the size of the system and its intended operating temperature. Chilling systems are more tolerant of water than freezers, and industrial systems with large liquid receivers are likely to be more tolerant than small direct-expansion (DX) circuits. It is imperative that the CO2 is specified with a suitable water content. Refrigerant grade, with a content less than 5 ppm, is suitable for small commercial systems; larger plant may use cryogenic grade, with a content less than 20 ppm. The content should be certified by the vendor and tested on site before installing in the system. On small systems, it may also be appropriate to charge through a filter-drier.

SYSTEM SAFETY Safety is an important factor in the design of every refrigeration system, and is one of the main reasons why carbon dioxide is gaining acceptance as a refrigerant of the future. CO2 is a natural refrigerant and considered environmentally safe. As a refrigerant, it is not without potential risks, but they are substantially smaller than those of other refrigerants. It is a slightly toxic, odorless, colorless gas with a slightly pungent, acid taste. Carbon dioxide is a small but important constituent of air. CO2 will not burn or support combustion. An atmosphere containing of more than 10% CO2 will extinguish an open flame. Mechanical failure in refrigeration equipment and piping can course a rapid increase in concentration levels of CO2. When inhaled at elevated concentrations, carbon dioxide may produce mild narcotic effects, stimulation of the respiratory centre, and asphyxiation, depending on concentration present. In the United States, the Occupational Safety and Health Administration (OSHA) limits the permissible exposure limit (PEL) time weighted average (TWA) concentration that must not be exceed during any 8 h per day, 40 h per week, to 5000 ppm. The OSHA shortterm exposure limit (STEL), a 15 min TWA exposure that should not be exceeded, is 30 000 ppm. In other countries (e.g., the United Kingdom), the STEL is lower, at 15 000 ppm. At atmospheric pressure, carbon dioxide is a solid, which sublimes to vapor at –56.6°C. All parts of a charged CO2 refrigerating system are above atmospheric pressure. Do not attempt to break piping joints or to remove valves or components without first

Carbon Dioxide Piping Materials When selecting piping material for CO2 refrigeration systems, the operating pressure and temperature requirements must be understood. Suitable piping materials may include copper, carbon steel, stainless steel, and aluminum. Many transcritical systems standardize on brazed air-conditioning and refrigeration (ACR) copper piping for the low-pressure side of the system, because of its availability. For pressures above 4.1 MPa, the annealing effect of brazing can weaken copper pipe, so pipework should be welded steel. Alternatively, cold-formed mechanical permanent joints can be used with copper pipe if the pipe and fittings are suitably pressure rated. Small-diameter copper tubing meets the requirement pressure ratings. The allowable internal pressure for copper tubing in service is based on a formula used in ASME Standard B31 and ASTM Standard 280: 2St m p = -----------------------------D – 0.08t m

(2)

where p = allowable pressure S = allowable stress [i.e., allowable design strength for continuous long-term service, from ASME (2007)] tm = wall thickness D = outside diameter

Carbon Steel Piping for CO2 Low-temperature seamless carbon steel pipe (ASTM Standard A333) Grade 6 is suited for conditions within refrigeration systems. Alternatively a number of common stainless steel alloys provide adequate low temperature properties. Stainless steel, aluminum, and carbon steel piping require qualified welders for the piping installation.

Pipe Sizing For the same pressure drop, CO2 has a corresponding temperature penalty 5 to 10 times smaller than ammonia and R-134a have (Figure 11). For a large system with an inherently large pressure drop, the temperature penalty with CO2 is substantially less than the same pressure drop using another refrigerant. Because of CO2’s physical properties (particularly density), the vapor side of the system is much smaller than in a typical ammonia system, but the liquid side is similar or even larger because CO2’s lower latent heat requires more mass flow (see Table 3). The primary method of sizing CO2 pipe is to define the allowable temperature loss that the system can handle, convert that to pressure loss, then size the

3.8

2014 ASHRAE Handbook—Refrigeration (SI) of the wet suction volume than with ammonia, so there is a significant advantage in reducing the circulating rate. Typically 2:1 can be used for a cold store, whereas 4:1 would be preferred in this application for ammonia. Design of a recirculator vessel must consider liquid flow rates. When sizing pump flow rates, the pump manufacturer’s recommendations for liquid velocity should generally be followed: • NH3 and most hydrocarbons (HCs):